Stirling cycle machine

ABSTRACT

A Stirling cycle machine. The machine includes at least one rocking drive mechanism which includes: a rocking beam having a rocker pivot, at least one cylinder and at least one piston. The piston is housed within a respective cylinder and is capable of substantially linearly reciprocating within the respective cylinder. Also, the drive mechanism includes at least one coupling assembly having a proximal end and a distal end. The linear motion of the piston is converted to rotary motion of the rocking beam. Also, a crankcase housing the rocking beam and housing a first portion of the coupling assembly is included. The machine also includes a working space housing the at least one cylinder, the at least one piston and a second portion of the coupling assembly. An airlock is included between the workspace and the crankcase and a seal is included for sealing the workspace from the airlock and crankcase. A burner and burner control system is also included for heating the machine and controlling ignition and combustion in the burner.

CROSS REFERENCE TO RELATED APPLICATIONS

This application is a Continuation Application of U.S. patentapplication Ser. No. 14/319,214, filed Jun. 30, 2014 and entitledStirling Cycle Machine, now U.S. Pat. No. 9,797,340, issued Oct. 24,2017, which is a Continuation Application of U.S. patent applicationSer. No. 12/829,329, filed Jul. 1, 2010 and entitled Stirling CycleMachine, now U.S. Pat. No. 8,763,391, issued Jul. 1, 2014 which is aContinuation-In-Part Application of prior U.S. patent application Ser.No. 12/105,854, filed Apr. 18, 2008 and entitled Stirling Cycle Machine,which is now U.S. Pat. No. 8,474,256, issued Jul. 2, 2013, which claimsthe benefit of U.S. Provisional Application No. 60/925,818, filed Apr.23, 2007 and entitled Four Cylinder Stirling Engine and U.S. ProvisionalApplication No. 60/925,814 filed Apr. 23, 2007 and entitled Rocking BeamDrive, all of which are hereby incorporated herein by reference in theirentireties.

U.S. patent application Ser. No. 12/829,329 also claims the benefit ofU.S. Provisional Application No. 61/222,361 filed Jul. 1, 2009 andentitled Stirling Cycle Machine, which is hereby incorporated herein byreference in its entirety.

TECHNICAL FIELD

The present invention relates to machines and more particularly, to aStirling cycle machine and components thereof.

BACKGROUND INFORMATION

Many machines, such as internal combustion engines, external combustionengines, compressors, and other reciprocating machines, employ anarrangement of pistons and drive mechanisms to convert the linear motionof a reciprocating piston to rotary motion. In most applications, thepistons are housed in a cylinder. A common problem encountered with suchmachines is that of friction generated by a sliding piston resultingfrom misalignment of the piston in the cylinder and lateral forcesexerted on the piston by linkage of the piston to a rotating crankshaft.These increased side loads increase engine noise, increase piston wear,and decrease the efficiency and life of the engine. Additionally,because of the side loads, the drive requires more power to overcomethese frictional forces, thus reducing the efficiency of the machine.

Improvements have been made on drive mechanisms in an attempt to reducethese side loads, however, many of the improvements have resulted inheavier and bulkier machines.

Accordingly, there is a need for practical machines with minimal sideloads on pistons.

SUMMARY

In accordance with one aspect of the present invention, a rocking beamdrive mechanism for a machine is disclosed. The drive mechanism includesa rocking beam having a rocker pivot, at least one cylinder and at leastone piston. The piston is housed within a respective cylinder. Thepiston is capable of substantially linearly reciprocating within therespective cylinder. Also, the drive mechanism includes at least onecoupling assembly having a proximal end and a distal end. The proximalend is connected to the piston and the distal end is connected to therocking beam by an end pivot. The linear motion of the piston isconverted to rotary motion of the rocking beam.

Some embodiments of this aspect of the present invention include one ormore of the following: where the rocking beam is coupled to a crankshaftby way of a connecting rod. In this embodiment, the rotary motion of therocking beam is transferred to the crankshaft. Also, where the cylindermay further include a closed end and an open end. The open end furtherincludes a linear bearing connected to the cylinder. The linear bearingincludes an opening to accommodate the coupling assembly. Also, wherethe coupling assembly further includes a piston rod and a link rod. Thepiston rod and link rod are coupled together by a coupling means. Thecoupling means is located beneath the linear bearing. Also, where thedrive mechanism also includes a seal, where the seal is sealablyconnected to the piston rod. Also, where the seal is a rollingdiaphragm. Also, in some embodiments, the coupling means is a flexiblejoint. In some embodiments, the coupling means is a roller bearing. Insome embodiments, the coupling means is a hinge. In some embodiments,the coupling means is a flexure. In some embodiments, the coupling meansis a journal bearing joint.

In accordance with another aspect of the present invention, a Stirlingcycle machine is disclosed. The machine includes at least one rockingdrive mechanism where the rocking drive mechanism includes: a rockingbeam having a rocker pivot, at least one cylinder and at least onepiston. The piston is housed within a respective cylinder. The piston iscapable of substantially linearly reciprocating within the respectivecylinder. Also, the drive mechanism includes at least one couplingassembly having a proximal end and a distal end. The proximal end isconnected to the piston and the distal end is connected to the rockingbeam by an end pivot. The linear motion of the piston is converted torotary motion of the rocking beam. Also, a crankcase housing the rockingbeam and housing a first portion of the coupling assembly is included. Acrankshaft coupled to the rocking beam by way of a connecting rod isalso included. The rotary motion of the rocking beam is transferred tothe crankshaft. The machine also includes a working space housing the atleast one cylinder, the at least one piston and a second portion of thecoupling assembly. A seal is included for sealing the workspace from thecrankcase.

Some embodiments of this aspect of the present invention include one ormore of the following: where the seal is a rolling diaphragm. Also, thecylinder may further include a closed end and an open end. The open endfurther includes a linear bearing connected to the cylinder. The linearbearing includes an opening to accommodate the coupling assembly. Also,where the coupling assembly further includes a piston rod and a linkrod. The piston rod and link rod are coupled together by a couplingmeans. The coupling means may be located beneath the linear bearing.Also, the machine may also include a lubricating fluid pump in thecrankcase. In some embodiments, the lubricating fluid pump is amechanical lubricating fluid pump driven by a pump drive assembly, thepump drive assembly being connected to and driven by the crankshaft. Insome embodiments, the lubricating fluid pump is an electric lubricatingfluid pump. The machine may also include a motor connected to thecrankshaft. The machine may also include a generator connected to thecrankshaft.

In accordance with another aspect of the present invention, a Stirlingcycle machine is disclosed. The machine includes at least two rockingdrive mechanisms. The rocking drive mechanisms each include a rockingbeam having a rocker pivot, two cylinders, and two pistons. The pistonseach housed within a respective cylinder. The pistons are capable ofsubstantially linearly reciprocating within the respective cylinder.Also, the drive mechanisms include two coupling assemblies having aproximal end and a distal end, the proximal end being connected to thepiston and the distal end being connected to the rocking beam by an endpivot. The linear motion of the piston is converted to rotary motion ofthe rocking beam. The machine also includes a crankcase housing therocking beam and housing a first portion of the coupling assemblies.Also, a crankshaft coupled to the rocking beam by way of a connectingrod. The rotary motion of the rocking beam is transferred to thecrankshaft. The machine also includes a lubricating fluid pump in thecrankcase for pumping lubricating fluid to lubricate the crankshaft andthe rocking beam and the first portion of the coupling assemblies. Also,a working space housing the cylinders, the pistons and the secondportion of the coupling assemblies. A rolling diaphragm for sealing theworkspace from the crankcase is also included.

Some embodiments of this aspect of the present invention include one ormore of the following: where the cylinder may further include a closedend and an open end. The open end further includes a linear bearingconnected to the cylinder. The linear bearing includes an opening toaccommodate the coupling assembly. Also, where the coupling assemblyfurther includes a piston rod and a link rod. The piston rod and linkrod are coupled together by a coupling means. The coupling means may belocated beneath the linear bearing. Also, where the coupling means is aflexible joint. In some embodiments, also disclosed is where thecoupling means is a roller bearing.

Other embodiments of this aspect of the present invention relate to oneor more of a rocking beam drive mechanism for a machine comprising arocking beam having a rocker pivot, at least one cylinder, at least onepiston, the piston housed within a respective cylinder whereby thepiston is capable of substantially linearly reciprocating within therespective cylinder, and at least one coupling assembly having aproximal end and a distal end, the proximal end being connected to thepiston and the distal end being connected to the rocking beam by an endpivot and linear motion of the piston is converted to rotary motion ofthe rocking beam.

A still further embodiment of the invention relate to one or moreembodiments of a Stirling cycle machine comprising at least one rockingdrive mechanism comprising a rocking beam having a rocker pivot, atleast one cylinder, at least one piston, the piston housed within arespective cylinder whereby the piston is capable of substantiallylinearly reciprocating within the respective cylinder and at least onecoupling assembly having a proximal end and a distal end, the proximalend being connected to the piston and the distal end being connected tothe rocking beam by an end pivot, whereby linear motion of the piston isconverted to rotary motion of the rocking beam, a crankcase housing therocking beam and housing a first portion of the coupling assembly, acrankshaft coupled to the rocking beam by way of a connecting rod,whereby the rotary motion of the rocking beam is transferred to thecrankshaft, a working space housing the at least one cylinder, the atleast one piston and a second portion of the coupling assembly, and anairlock space separating the crankcase and the working space formaintaining a pressure differential between the crankcase housing andthe working space housing.

A still further embodiment of the invention relate to one or moreembodiments of an external combustion engine comprising at least tworocking drive mechanisms comprising a rocking beam having a rockerpivot, at least two cylinders, at least two pistons, the pistons eachhoused within a respective cylinder whereby the pistons are capable ofsubstantially linearly reciprocating within the respective cylinder andtwo coupling assemblies having a proximal end and a distal end, theproximal end being connected to the piston and the distal end beingconnected to the rocking beam by an end pivot, whereby linear motion ofthe piston is converted to rotary motion of the rocking beam, acrankcase housing the rocking beam and housing a first portion of thecoupling assemblies, a crankshaft coupled to the rocking beam by way ofa connecting rod, whereby the rotary motion of the rocking beam istransferred to the crankshaft, a lubricating fluid pump in the crankcasefor pumping lubricating fluid to lubricate the crankshaft and therocking beam and the first portion of the coupling assemblies, a workingspace housing the cylinders, the pistons and the second portion of thecoupling assemblies, an airlock space separating the crankcase and theworking space for maintaining a pressure differential between thecrankcase housing and the working space housing, a heating elementcomprising a burner having at least one burner head for igniting andmaintaining a heating flame in a combustion chamber adjacent the atleast one heater head, and an electronic control unit managing theheating element according to operational data of the engine obtainedfrom at least one of the rocking drive mechanisms, lubricating fluidpump, the crankcase, the working space, crankshaft, heating element andthe airlock.

These aspects of the invention are not meant to be exclusive and otherfeatures, aspects, and advantages of the present invention will bereadily apparent to those of ordinary skill in the art when read inconjunction with the appended claims and accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

These and other features and advantages of the present invention will bebetter understood by reading the following detailed description, takentogether with the drawings wherein:

FIGS. 1A-1E depict the principle of operation of a prior art Stirlingcycle machine;

FIG. 2 shows a view of a rocking beam drive in accordance with oneembodiment;

FIG. 3 shows a view of a rocking beam drive in accordance with oneembodiment;

FIG. 4 shows a view of an engine in accordance with one embodiment;

FIGS. 5A-5D depicts various views of a rocking beam drive in accordancewith one embodiment;

FIG. 6 shows a bearing style rod connector in accordance with oneembodiment;

FIGS. 7A-7B show a flexure in accordance with one embodiment;

FIG. 8 shows a four cylinder double rocking beam drive arrangement inaccordance with one embodiment;

FIG. 9 shows a cross section of a crankshaft in accordance with oneembodiment;

FIG. 10A shows a view of an engine in accordance with one embodiment;

FIG. 10B shows a crankshaft coupling in accordance with one embodiment;

FIG. 10C shows a view of a sleeve rotor in accordance with oneembodiment;

FIG. 10D shows a view of a crankshaft in accordance with one embodiment;

FIG. 10E is a cross section of the sleeve rotor and spline shaft inaccordance with one embodiment;

FIG. 10F is a cross section of the crankshaft and the spline shaft inaccordance with one embodiment;

FIG. 10G are various views a sleeve rotor, crankshaft and spline shaftin accordance with one embodiment;

FIG. 11 shows the operation of pistons of an engine in accordance withone embodiment;

FIG. 12A shows an unwrapped schematic view of a working space andcylinders in accordance with one embodiment;

FIG. 12B shows a schematic view of a cylinder, heater head, andregenerator in accordance with one embodiment;

FIG. 12C shows a view of a cylinder head in accordance with oneembodiment;

FIG. 13A shows a view of a rolling diaphragm, along with supporting topseal piston and bottom seal piston, in accordance with one embodiment;

FIG. 13B shows an exploded view of a rocking beam driven engine inaccordance with one embodiment;

FIG. 13C shows a view of a cylinder, heater head, regenerator, androlling diaphragm, in accordance with one embodiment;

FIGS. 13D-13E show various views of a rolling diaphragm duringoperation, in accordance with one embodiment;

FIG. 13F shows an unwrapped schematic view of a working space andcylinders in accordance with one embodiment;

FIG. 13G shows a view of an external combustion engine in accordancewith one;

FIGS. 14A-14E show views of various embodiments of a rolling diaphragm;

FIG. 15A shows a view of a metal bellows and accompanying piston rod andpistons in accordance with one embodiment;

FIGS. 15B-15D show views of metal bellows diaphragms, in accordance withone embodiment;

FIGS. 15E-15G show a view of metal bellows in accordance with variousembodiments;

FIG. 15H shows a schematic of a rolling diaphragm identifying variousload regions;

FIG. 15I shows a schematic of the rolling diaphragm identifying theconvolution region;

FIG. 16 shows a view of a piston and piston seal in accordance with oneembodiment;

FIG. 17 shows a view of a piston rod and piston rod seal in accordancewith one embodiment;

FIG. 18A shows a view of a piston seal backing ring in accordance withone embodiment;

FIG. 18B shows a pressure diagram for a backing ring in accordance withone embodiment;

FIGS. 18C and 18D show a piston seal in accordance with one embodiment;

FIGS. 18E and 18F show a piston rod seal in accordance with oneembodiment;

FIG. 19A shows a view of a piston seal backing ring in accordance withone embodiment;

FIG. 19B shows a pressure diagram for a piston seal backing ring inaccordance with one embodiment;

FIG. 20A shows a view of a piston rod seal backing ring in accordancewith one embodiment;

FIG. 20B shows a pressure diagram for a piston rod seal backing ring inaccordance with one embodiment;

FIG. 21 shows views of a piston guide ring in accordance with oneembodiment;

FIG. 22 shows an unwrapped schematic illustration of a working space andcylinders in accordance with one embodiment;

FIG. 23A shows a view of an engine in accordance with one embodiment;

FIG. 23B shows a view of an engine in accordance with one embodiment;

FIG. 24 shows a view of a crankshaft in accordance with one embodiment;

FIGS. 25A-25C show various configurations of pump drives in accordancewith various embodiments;

FIG. 26A show various views of an oil pump in accordance with oneembodiment;

FIG. 26B shows a view of an engine in accordance with one embodiment;

FIG. 26C shows another view of the engine depicted in FIG. 26B;

FIGS. 27A and 27B show views of an engine in accordance with oneembodiment;

FIG. 27C shows a view of a coupling joint in accordance with oneembodiment;

FIG. 27D shows a view of a crankshaft and spline shaft of an engine inaccordance with one embodiment;

FIG. 28 shows a view of a heater exchanger and burner for an engine inaccordance with one embodiment;

FIG. 29 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 30 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 31 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 32 shows a view of heater tubes of a heat exchanger in accordancewith one embodiment;

FIG. 33 shows a view of heater tubes of a heat exchanger in accordancewith one embodiment;

FIG. 34 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 35 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 36 shows a view of a heater head of an engine in accordance withone embodiment;

FIG. 37 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 38 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 39 shows a portion of a cross section of a tube heat exchanger inaccordance with one embodiment;

FIG. 40 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 41 shows a portion of a cross section of a tube heat exchanger inaccordance with one embodiment;

FIG. 42 shows a view of a heater head of an engine in accordance withone embodiment;

FIG. 43A shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 43B shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 44A shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 44B shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 45A shows view of a tube heat exchanger in accordance with oneembodiment;

FIG. 45B shows a view of a tube heat exchanger in accordance with oneembodiment;

FIGS. 46A-46D show various configurations of a tube heat exchanger inaccordance with various embodiments;

FIG. 47 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 48 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 49 shows a view of a heater head of an engine in accordance withone embodiment;

FIG. 50 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIGS. 51A and 51B show views of heat exchangers of an engine inaccordance with various embodiments;

FIGS. 52A-52C show various views of a heat exchanger in accordance withone embodiment;

FIG. 52D shows a view of a heat exchanger in accordance with oneembodiment;

FIGS. 53A and 53B show views of a heat exchanger in accordance with oneembodiment;

FIG. 53C shows a view of a heat exchanger of an engine in accordancewith one embodiment;

FIGS. 53D-53F show views of a heat exchanger of an engine in accordancewith one embodiment;

FIGS. 54A and 54B show views of a heat exchanger of an engine inaccordance with one embodiment;

FIGS. 55A-55D show various views of a heat exchanger in accordance withone embodiment;

FIGS. 56A-56C show various configurations of a heat exchanger inaccordance with various embodiments;

FIGS. 57A and 57B show various diagrams depicting physical properties ofa heat exchanger in accordance with one embodiment;

FIG. 58 shows a view of a heater head in accordance with one embodiment;

FIG. 59 shows a view of a heater head in accordance with one embodiment;

FIGS. 60A and 60B show views of a heater head in accordance with oneembodiment;

FIGS. 61A and 61B show views of a heater head in accordance with oneembodiment;

FIGS. 62A and 62B show views of a heater head in accordance with oneembodiment;

FIG. 62C shows a views of a heater head in accordance with oneembodiment;

FIG. 62D shows a view of a heater head in accordance with oneembodiment;

FIG. 62E shows a view of a heater head in accordance with oneembodiment;

FIGS. 63A and 63B show a regenerator of a Stirling cycle engine inaccordance with one embodiment;

FIGS. 64A-64E show various configurations of a regenerator of a Stirlingcycle engine in accordance with various embodiments;

FIGS. 65A-65G show various views of an engine in accordance with severalembodiments;

FIGS. 66A and 66B show views of a cooler for an engine in accordancewith some embodiments;

FIGS. 67A-67B show views of a cooler for an engine in accordance withone embodiment;

FIG. 67C-67D show views of a cooler for an engine in accordance with oneembodiment;

FIG. 67E shows a view of the embodiment of a cooler for an enginedepicted in FIG. 67A;

FIG. 68 shows a view of an intake manifold for an engine in accordancewith one embodiment;

FIGS. 69A and 69B show various views of an intake manifold for an enginein accordance with one embodiment;

FIG. 70 shows a view of a heater head of an engine in accordance withyet another embodiment of the invention;

FIGS. 71A and 71B show views of a burner of an engine in accordance withone embodiment;

FIG. 72 is a gaseous fuel burner coupled to a Stirling cycle engine,where the ejector is a venturi, according to one embodiment;

FIG. 73A is the burner of FIG. 72 showing the air and fuel flow paths;

FIG. 73B is a graphical representation of the pressure across theburner;

FIG. 74 shows a view of a venturi as shown in the burner of FIG. 72;

FIGS. 75 and 75A are embodiments of the venturi in FIG. 72;

FIG. 75B shows a schematic of a multiple fuel system with multiple fuelrestrictions and valves;

FIG. 76 shows a schematic of an embodiment of the burner with automatedfuel control for variable fuel properties;

FIG. 77 shows a schematic of another embodiment of the burner withtemperature sensor and engine speed control loop;

FIG. 78 shows a schematic of yet another embodiment of the burner withtemperature sensor and oxygen sensor control loop;

FIG. 79 shows an alternative embodiment of the ejector wherein the fuelis fed directly into the ejector;

FIG. 80 is a block diagram showing a system for controlling apressurized combustion chamber of an engine according to an embodiment;

FIG. 81 shows a piston pump according to one embodiment;

FIG. 82 shows an alternating current waveform suitable for driving thepiston pump of FIG. 81;

FIG. 83 shows a pulse-width-modulated direct current waveform suitablefor driving the piston pump of FIG. 81, according to one embodiment;

FIG. 84 is schematic diagram of a diaphragm pump according to oneembodiment;

FIG. 85 is a schematic diagram of a center-tapped coil for a diaphragmpump according to one embodiment;

FIGS. 86A and 86B shows pulse-width-modulated direct current waveformssuitable for driving the center-tapped coil of FIG. 85, according tosome embodiments;

FIGS. 87A-87D show embodiments of including a filter between the fuelpump and combustion chamber;

FIG. 88 shows a view of an engine in accordance with one embodiment;

FIGS. 89A-89C show views of a burner for an engine in accordance withvarious embodiments;

FIG. 90 shows a view of an engine with multiple burners in accordancewith yet another embodiment of the invention;

FIGS. 91A and 91B show views of multiple burners for an engine inaccordance with various embodiments;

FIG. 91C shows a view of a tube heater head in accordance with oneembodiment;

FIG. 91D shows a cross section of the tube heater head depicted in FIG.91C;

FIG. 92 shows a cross section of an engine in accordance with oneembodiment;

FIG. 93A-B shows views of a heater tube with an insert;

FIG. 94A-B shows an embodiment of a helical heater tube insert;

FIG. 95A-B show cross sectional views of two embodiments of a heatertube insert in a heater tube.

FIGS. 96A and 96B show a cross-sectional view of a Stirling cyclemachine having an inverted rocking beam design in accordance with oneembodiment;

FIGS. 96C-96E show various views of a piston and piston rod assembly inaccordance with one embodiment;

FIG. 96F shows a view of a heater tube in accordance with oneembodiment;

FIG. 97A shows a view of an embodiment of the rocking beam with a conrodbearing ratio of 1.6;

FIG. 97B shows a view of an embodiment of the rocking beam with a conrodbearing ratio of 1.0;

FIG. 98A shows an oil pump according to one embodiment;

FIG. 98B shows a Gerotor displacement pumping unit according to oneembodiment;

FIG. 99A shows a test rig assembly of a high pressure rod seal accordingto one embodiment;

FIG. 99B shows an embodiment of a high pressure rod seal;

FIG. 100A shows another embodiment of a high pressure rod seal includinga spring energized lip seal;

FIG. 100B is a hydraulic high pressure piston rod seal set inside therod seal cavity of a test rig according to one embodiment;

FIGS. 101A and 101B show views of a rolling diaphragm in accordance withone embodiment;

FIGS. 102A and 102B show views of a rolling diaphragm in accordance withanother embodiment;

FIG. 103 shows a view of a double bellows system in accordance with oneembodiment;

FIGS. 104A and 104B show views of an airlock pressure regulation systemin accordance with one embodiment;

FIG. 104C shows a bidirectional regulator according to one embodiment;

FIGS. 104D-104H show various positions of a spool valve in abidirectional regulator in accordance with various embodiments;

FIG. 1041 shows a view of an airlock pressure regulation system inaccordance with one embodiment;

FIG. 105 shows a view of an airlock pressure regulation system inaccordance with one embodiment;

FIG. 106 shows a view of a mechanical pump for regulating airlockpressure in accordance with one embodiment;

FIGS. 107A and 107B show views of a heat exchanger in accordance withone embodiment;

FIGS. 108A and 108B show views of a rocking beam mechanism in accordancewith one embodiment;

FIGS. 109A and 109B show views of a horizontally supported Stirlingcycle engine in accordance with one embodiment;

FIGS. 110A and 110B show views of a tube-in-tube heat exchangeraccording to one embodiment;

FIGS. 111A and 111B show views of a pull-cord start system in accordancewith one embodiment;

FIGS. 111C and 111D show views of an electric starter motor inaccordance with one embodiment;

FIGS. 112-117 show various views of a burner in accordance with oneembodiment;

FIGS. 118A-118C show various configurations of a fletching in accordancewith various embodiments;

FIGS. 118D and 119 show views of a burner in accordance with oneembodiment;

FIG. 120 is a diagram of a control burner scheme in accordance with oneembodiment;

FIGS. 121 and 122 show views of a burner in accordance with oneembodiment;

FIG. 123 shows a diagram of a partial combustion cooling process inaccordance with one embodiment;

FIG. 124 shows a diagram of a trim air cooling process in accordancewith one embodiment;

FIG. 125A shows a cutaway view of a restricting flow apparatus inaccordance with one embodiment;

FIG. 125B shows an isometric view of a burner housing in accordance withthe restricting flow apparatus in accordance with one embodiment;

FIG. 126 shows a control scheme in accordance with one embodiment;

FIG. 127 is a temperature distribution chart in accordance with oneembodiment of the restrictive flow apparatus; and

FIG. 128 is an example of a result graph in accordance with oneembodiment of the restrictive flow apparatus.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

Stirling cycle machines, including engines and refrigerators, have along technological heritage, described in detail in Walker, StirlingEngines, Oxford University Press (1980), incorporated herein byreference. The principle underlying the Stirling cycle engine is themechanical realization of the Stirling thermodynamic cycle:isovolumetric heating of a gas within a cylinder, isothermal expansionof the gas (during which work is performed by driving a piston),isovolumetric cooling, and isothermal compression. Additional backgroundregarding aspects of Stirling cycle machines and improvements thereto isdiscussed in Hargreaves, The Phillips Stirling Engine (Elsevier,Amsterdam, 1991), which is herein incorporated by reference.

The principle of operation of a Stirling cycle machine is readilydescribed with reference to FIGS. 1A-1E, wherein identical numerals areused to identify the same or similar parts. Many mechanical layouts ofStirling cycle machines are known in the art, and the particularStirling cycle machine designated generally by numeral 10 is shownmerely for illustrative purposes. In FIGS. 1A to 1D, piston 12 and adisplacer 14 move in phased reciprocating motion within the cylinders 16which, in some embodiments of the Stirling cycle machine, may be asingle cylinder, but in other embodiments, may include greater than asingle cylinder. A working fluid contained within cylinders 16 isconstrained by seals from escaping around piston 12 and displacer 14.The working fluid is chosen for its thermodynamic properties, asdiscussed in the description below, and is typically helium at apressure of several atmospheres, however, any gas, including any inertgas, may be used, including, but not limited to, hydrogen, argon, neon,nitrogen, air and any mixtures thereof. The position of the displacer 14governs whether the working fluid is in contact with the hot interface18 or the cold interface 20, corresponding, respectively, to theinterfaces at which heat is supplied to and extracted from the workingfluid. The supply and extraction of heat is discussed in further detailbelow. The volume of working fluid governed by the position of thepiston 12 is referred to as the compression space 22.

During the first phase of the Stirling cycle, the starting condition ofwhich is depicted in FIG. 1A, the piston 12 compresses the fluid in thecompression space 22. The compression occurs at a substantially constanttemperature because heat is extracted from the fluid to the ambientenvironment. The condition of the Stirling cycle machine 10 aftercompression is depicted in FIG. 1B. During the second phase of thecycle, the displacer 14 moves in the direction of the cold interface 20,with the working fluid displaced from the region of the cold interface20 to the region of the hot interface 18. This phase may be referred toas the transfer phase. At the end of the transfer phase, the fluid is ata higher pressure since the working fluid has been heated at constantvolume. The increased pressure is depicted symbolically in FIG. 1C bythe reading of the pressure gauge 24.

During the third phase (the expansion stroke) of the Stirling cyclemachine, the volume of the compression space 22 increases as heat isdrawn in from outside the Stirling cycle machine 10, thereby convertingheat to work. In practice, heat is provided to the fluid by means of aheater head (not shown) which is discussed in greater detail in thedescription below. At the end of the expansion phase, the compressionspace 22 is full of cold fluid, as depicted in FIG. 1D. During thefourth phase of the Stirling cycle machine 10, fluid is transferred fromthe region of the hot interface 18 to the region of the cold interface20 by motion of the displacer 14 in the opposing sense. At the end ofthis second transfer phase, the fluid fills the compression space 22 andcold interface 20, as depicted in FIG. 1A, and is ready for a repetitionof the compression phase. The Stirling cycle is depicted in a P-V(pressure-volume) diagram as shown in FIG. 1E.

Additionally, on passing from the region of the hot interface 18 to theregion of the cold interface 20. In some embodiments, the fluid may passthrough a regenerator (shown as 408 in FIG. 4). A regenerator is amatrix of material having a large ratio of surface area to volume whichserves to absorb heat from the fluid when it enters from the region ofthe hot interface 18 and to heat the fluid when it passes from theregion of the cold interface 20.

Stirling cycle machines have not generally been used in practicalapplications due to several daunting challenges to their development.These involve practical considerations such as efficiency and lifetime.Accordingly, there is a need for more Stirling cycle machines withminimal side loads on pistons, increased efficiency and lifetime.

The principle of operation of a Stirling cycle machine or Stirlingengine is further discussed in detail in U.S. Pat. No. 6,381,958, issuedMay 7, 2002, to Kamen et al., which is herein incorporated by referencein its entirety.

Rocking Beam Drive

Referring now to FIGS. 2-4, embodiments of a Stirling cycle machine,according to one embodiment, are shown in cross-section. The engineembodiment is designated generally by numeral 300. While the Stirlingcycle machine will be described generally with reference to the Stirlingengine 300 embodiments shown in FIGS. 2-4, it is to be understood thatmany types of machines and engines, including but not limited torefrigerators and compressors may similarly benefit from variousembodiments and improvements which are described herein, including butnot limited to, external combustion engines and internal combustionengines.

FIG. 2 depicts a cross-section of an embodiment of a rocking beam drivemechanism 200 (the term “rocking beam drive” is used synonymously withthe term “rocking beam drive mechanism”) for an engine, such as aStirling engine, having linearly reciprocating pistons 202 and 204housed within cylinders 206 and 208, respectively. The cylinders includelinear bearings 220. Rocking beam drive 200 converts linear motions ofpistons 202 and 204 into the rotary motion of a crankshaft 214. Rockingbeam drive 200 has a rocking beam 216, rocker pivot 218, a firstcoupling assembly 210, and a second coupling assembly 212. Pistons 202and 204 are coupled to rocking beam drive 200, respectively, via firstcoupling assembly 210 and second coupling assembly 212. The rocking beamdrive is coupled to crankshaft 214 via a connecting rod 222.

In some embodiments, the rocking beam and a first portion of thecoupling assembly may be located in a crankcase, while the cylinders,pistons and a second portion of the coupling assembly is located in aworkspace.

In FIG. 4 a crankcase 400 most of the rocking beam drive 200 ispositioned below the cylinder housing 402. Crankcase 400 is a space topermit operation of rocking beam drive 200 having a crankshaft 214,rocking beam 216, linear bearings 220, a connecting rod 222, andcoupling assemblies 210 and 212. Crankcase 400 intersects cylinders 206and 208 transverse to the plane of the axes of pistons 202 and 204.Pistons 202 and 204 reciprocate in respective cylinders 206 and 208, asalso shown in FIG. 2. Cylinders 206 and 208 extend above crankshafthousing 400. Crankshaft 214 is mounted in crankcase 400 below cylinders206 and 208.

FIG. 2 shows one embodiment of rocking beam drive 200. Couplingassemblies 210 and 212 extend from pistons 202 and 204, respectively, toconnect pistons 202 and 204 to rocking beam 216. Coupling assembly 212for piston 204, in some embodiments, may comprise a piston rod 224 and alink rod 226. Coupling assembly 210 for piston 202, in some embodiments,may comprise a piston rod 228 and a link rod 230. Piston 204 operates inthe cylinder 208 vertically and is connected by the coupling assembly212 to the end pivot 232 of the rocking beam 216. The cylinder 208provides guidance for the longitudinal motion of piston 204. The pistonrod 224 of the coupling assembly 212 attached to the lower portion ofpiston 204 is driven axially by its link rod 226 in a substantiallylinear reciprocating path along the axis of the cylinder 208. The distalend of piston rod 224 and the proximate end of link rod 226, in someembodiments, may be jointly hinged via a coupling means 234. Thecoupling means 234, may be any coupling means known in the art,including but not limited to, a flexible joint, roller bearing element,hinge, journal bearing joint (shown as 600 in FIG. 6), and flexure(shown as 700 in FIGS. 7A and 7B). The distal end of the link rod 226may be coupled to one end pivot 232 of rocking beam 216, which ispositioned vertically and perpendicularly under the proximate end of thelink rod 226. A stationary linear bearing 220 may be positioned alongcoupling assembly 212 to further ensure substantially linearlongitudinal motion of the piston rod 224 and thus ensuringsubstantially linear longitudinal motion of the piston 204. In anexemplary embodiment, link rod 226 does not pass through linear bearing220. This ensures, among other things, that piston rod 224 retains asubstantially linear and longitudinal motion.

In the exemplary embodiment, the link rods may be made from aluminum,and the piston rods and connecting rod are made from D2 Tool Steel.Alternatively, the link rods, piston rods, connecting rods, and rockingbeam may be made from 4340 steel. Other materials may be used for thecomponents of the rocking beam drive, including, but not limited to,titanium, aluminum, steel or cast iron. In some embodiments, the fatiguestrength of the material being used is above the actual load experiencedby the components during operation.

Still referring to FIGS. 2-4, piston 202 operates vertically in thecylinder 206 and is connected by the coupling assembly 210 to the endpivot 236 of the rocking beam 216. The cylinder 206 serves, amongstother functions, to provide guidance for longitudinal motion of piston202. The piston rod 228 of the coupling assembly 210 is attached to thelower portion of piston 202 and is driven axially by its link rod 230 ina substantially linear reciprocating path along the axis of the cylinder206. The distal end of the piston rod 228 and the proximate end of thelink rod 230, in some embodiments, is jointly hinged via a couplingmeans 238. The coupling means 238, in various embodiments may include,but are not limited to, a flexure (shown as 700 in FIGS. 7A and 7B,roller bearing element, hinge, journal bearing (shown as 600 in FIG. 6),or coupling means as known in the art. The distal end of the link rod230, in some embodiments, may be coupled to one end pivot 236 of rockingbeam 216, which is positioned vertically and perpendicularly under theproximate end of link rod 230. A stationary linear bearing 220 may bepositioned along coupling assembly 210 to further ensure linearlongitudinal motion of the piston rod 228 and thus ensuring linearlongitudinal motion of the piston 202. In an exemplary embodiment, linkrod 230 does not pass through linear bearing 220 to ensure that pistonrod 228 retains a substantially linear and longitudinal motion.

The coupling assemblies 210 and 212 change the alternating longitudinalmotion of respective pistons 202 and 204 to oscillatory motion of therocking beam 216. The delivered oscillatory motion is changed to therotational motion of the crankshaft 214 by the connecting rod 222,wherein one end of the connecting rod 222 is rotatably coupled to aconnecting pivot 240 positioned between an end pivot 232 and a rockerpivot 218 in the rocking beam 216, and another end of the connecting rod222 is rotatably coupled to crankpin 246. The rocker pivot 218 may bepositioned substantially at the midpoint between the end pivots 232 and236 and oscillatorily support the rocking beam 216 as a fulcrum, thusguiding the respective piston rods 224 and 228 to make sufficient linearmotion. In the exemplary embodiment, the crankshaft 214 is located abovethe rocking beam 216, but in other embodiments, the crankshaft 214 maybe positioned below the rocking beam 216 (as shown in FIGS. 5B and 5D)or in some embodiments, the crankshaft 214 is positioned to the side ofthe rocking beam 216, such that it still has a parallel axis to therocking beam 216.

Still referring to FIGS. 2-4, the rocking beam oscillates about therocker pivot 218, the end pivots 232 and 236 follow an arc path. Sincethe distal ends of the link rods 226 and 230 are connected to therocking beam 216 at pivots 232 and 236, the distal ends of the link rods226 and 230 also follow this arc path, resulting in an angular deviation242 and 244 from the longitudinal axis of motion of their respectivepistons 202 and 204. The coupling means 234 and 238 are configured suchthat any angular deviation 244 and 242 from the link rods 226 and 230experienced by the piston rods 224 and 228 is minimized. Essentially,the angular deviation 244 and 242 is absorbed by the coupling means 234and 238 so that the piston rods 224 and 228 maintain substantiallylinear longitudinal motion to reduce side loads on the pistons 204 and202. A stationary linear bearing 220 may also be placed inside thecylinder 208 or 206, or along coupling assemblies 212 or 210, to furtherabsorb any angular deviation 244 or 242 thus keeping the piston push rod224 or 228 and the piston 204 or 202 in linear motion along thelongitudinal axis of the piston 204 or 202.

Therefore, in view of reciprocating motion of pistons 202 and 204, it isnecessary to keep the motion of pistons 202 and 204 as close to linearas possible because the deviation 242 and 244 from longitudinal axis ofreciprocating motion of pistons 202 and 204 causes noise, reduction ofefficiency, increase of friction to the wall of cylinder, increase ofside-load, and low durability of the parts. The alignment of thecylinders 206 and 208 and the arrangement of crankshaft 214, piston rods224 and 228, link rods 226 and 230, and connecting rod 222, hence, mayinfluence on, amongst other things, the efficiency and/or the volume ofthe device. For the purpose of increasing the linearity of the pistonmotion as mentioned, the pistons (shown as 202 and 204 in FIGS. 2-4) arepreferably as close to the side of the respective cylinders 206 and 208as possible.

In another embodiment reducing angular deviation of link rods, link rods226 and 230 substantially linearly reciprocate along longitudinal axisof motion of respective pistons 204 and 202 to decrease the angulardeviation and thus to decrease the side load applied to each piston 204and 202. The angular deviation defines the deviation of the link rod 226or 230 from the longitudinal axis of the piston 204 or 202. Numerals 244and 242 designate the angular deviation of the link rods 226 and 230, asshown in FIG. 2. Therefore, the position of coupling assembly 212influences the angular displacement of the link rod 226, based on thelength of the distance between the end pivot 232 and the rocker pivot218 of the rocking beam 216. Thus, the position of the couplingassemblies may be such that the angular displacement of the link rod 226is reduced. For the link rod 230, the length of the coupling assembly210 also may be determined and placed to reduce the angular displacementof the link rod 230, based on the length of the distance between the endpivot 236 and the rocker pivot 218 of the rocking beam 216. Therefore,the length of the link rods 226 and 230, the length of couplingassemblies 212 and 210, and the length of the rocking beam 216 aresignificant parameters that greatly influence and/or determine theangular deviation of the link rods 226 and 230 as shown in FIG. 2.

The exemplary embodiment has a straight rocking beam 216 having the endpoints 232 and 236, the rocker pivot 218, and the connecting pivot 240along the same axis. However, in other embodiments, the rocking beam 216may be bent, such that pistons may be placed at angles to each other, asshown in FIGS. 5C and 5D.

Referring now to FIGS. 2-4 and FIGS. 7A-7B, in some embodiments of thecoupling assembly, the coupling assemblies 212 and 210, may include aflexible link rod that is axially stiff but flexible in the rocking beam216 plane of motion between link rods 226 and 230, and pistons 204 and202, respectively. In this embodiment, at least one portion, the flexure(shown as 700 in FIGS. 7A and 7B), of link rods 226 and 230 is elastic.The flexture 700 acts as a coupling means between the piston rod and thelink rod. The flexure 700 may absorb the crank-induced side loads of thepistons more effectively, thus allowing its respective piston tomaintain linear longitudinal movement inside the piston's cylinder. Thisflexure 700 allows small rotations in the plane of the rocking beam 216between the link rods 226 and 230 and pistons 204 or 202, respectively.Although depicted in this embodiment as flat, which increases theelasticity of the link rods 226 and 230, the flexure 700, in someembodiments, is not flat. The flexure 700 also may be constructed nearto the lower portion of the pistons or near to the distal end of thelink rods 226 and 230. The flexure 700, in one embodiment, may be madeof #D2 Tool Steel Hardened to 58-62 RC. In some embodiments, there maybe more than one flexure (not shown) on the link rod 226 or 230 toincrease the elasticity of the link rods.

In alternate embodiment, the axes of the pistons in each cylinderhousing may extend in different directions, as depicted in FIGS. 5C and5D. In the exemplary embodiment, the axes of the pistons in eachcylinder housing are substantially parallel and preferably substantiallyvertical, as depicted in FIGS. 2-4, and FIGS. 5A and 5B. FIGS. 5A-5Dinclude various embodiments of the rocking beam drive mechanismincluding like numbers as those shown and described with respect toFIGS. 2-4. It will be understood by those skilled in that art thatchanging the relative position of the connecting pivot 240 along therocking beam 216 will change the stroke of the pistons.

Accordingly, a change in the parameters of the relative position of theconnecting pivot 240 in the rocking beam 216 and the length of thepiston rods 224 and 228, link rods 230 and 226, rocking beam 216, andthe position of rocker pivot 218 will change the angular deviation ofthe link rods 226 and 230, the phasing of the pistons 204 and 202, andthe size of the device 300 in a variety of manner. Therefore, in variousembodiments, a wide range of piston phase angles and variable sizes ofthe engine may be chosen based on the modification of one or more ofthese parameters. In practice, the link rods 224 and 228 of theexemplary embodiment have substantially lateral movement within from−0.5 degree to +0.5 degree from the longitudinal axis of the pistons 204and 202. In various other embodiments, depending on the length of thelink rod, the angle may vary anywhere from approaching 0 degrees to 0.75degrees. However, in other embodiments, the angle may be higherincluding anywhere from approaching 0 to the approximately 20 degrees.As the link rod length increases, however, the crankcase/overall engineheight increases as well as the weight of the engine.

One feature of the exemplary embodiment is that each piston has its linkrod extending substantially to the attached piston rod so that it isformed as a coupling assembly. In one embodiment, the coupling assembly212 for the piston 204 includes a piston rod 224, a link rod 226, and acoupling means 234 as shown in FIG. 2. More specifically, one proximalend of piston rod 224 is attached to the lower portion of piston 204 andthe distal end piston rod 224 is connected to the proximate end of thelink rod 226 by the coupling means 234. The distal end of the link rod226 extends vertically to the end pivot 232 of the rocking beam 216. Asdescribed above, the coupling means 234 may be, but is not limited to, ajoint, hinge, coupling, or flexure or other means known in the art. Inthis embodiment, the ratio of the piston rod 224 and the link rod 226may determine the angular deviation of the link rod 226 as mentionedabove.

In one embodiment of the machine, an engine, such as a Stirling engine,employs more than one rocking beam drive on a crankshaft. Referring nowto FIG. 8, an unwrapped “four cylinder” rocking beam drive mechanism 800is shown. In this embodiment, the rocking beam drive mechanism has fourpistons 802, 804, 806, and 808 coupled to two rocking beam drives 810and 812. In the exemplary embodiment, rocking beam drive mechanism 800is used in a Stirling engine comprising at least four pistons 802, 804,806, and 808, positioned in a quadrilateral arrangement coupled to apair of rocking beam drives 810 and 812, wherein each rocking beam driveis connected to crankshaft 814. However, in other embodiments, theStirling cycle engine includes anywhere from 1-4 pistons, and in stillother embodiments, the Stirling cycle engine includes more than 4pistons. In some embodiments, rocking beam drives 810 and 812 aresubstantially similar to the rocking beam drives described above withrespect to FIGS. 2-4 (shown as 210 and 212 in FIGS. 2-4). Although inthis embodiment, the pistons are shown outside the cylinders, inpractice, the pistons would be inside cylinders.

Still referring to FIG. 8, in some embodiments, the rocking beam drivemechanism 800 has a single crankshaft 814 having a pair oflongitudinally spaced, radially and oppositely directed crank pins 816and 818 adapted for being journalled in a housing, and a pair of rockingbeam drives 810 and 812. Each rocking beam 820 and 822 is pivotallyconnected to rocker pivots 824 and 826, respectively, and to crankpins816 and 818, respectively. In the exemplary embodiment, rocking beams820 and 822 are coupled to a rocking beam shaft 828.

In some embodiments, a motor/generator may be connected to thecrankshaft in a working relationship. The motor may be located, in oneembodiment, between the rocking beam drives. In another embodiment, themotor may be positioned outboard. The term “motor/generator” is used tomean either a motor or a generator.

FIG. 9 shows one embodiment of crankshaft 814. Positioned on thecrankshaft is a motor/generator 900, such as a Permanent Magnetic (“PM”)generator. Motor/generator 900 may be positioned between, or inboard ofthe rocking beam drives (not shown, shown in FIG. 8 as 810 and 812), ormay be positioned outside, or outboard of, rocking beam drives 810 and812 at an end of crankshaft 814, as depicted by numeral 1000 in FIG.10A.

When motor/generator 900 is positioned between the rocking beam drives(not shown, shown in FIG. 8 as 810 and 812), the length ofmotor/generator 900 is limited to the distance between the rocking beamdrives. The diameter squared of motor/generator 900 is limited by thedistance between the crankshaft 814 and the rocking beam shaft 828.Because the capacity of motor/generator 900 is proportional to itsdiameter squared and length, these dimension limitations result in alimited-capacity “pancake” motor/generator 900 having relatively shortlength, and a relatively large diameter squared. The use of a “pancake”motor/generator 900 may reduce the overall dimension of the engine,however, the dimension limitations imposed by the inboard configurationresult in a motor/generator having limited capacity.

Placing motor/generator 900 between the rocking beam drives exposesmotor/generator 900 to heat generated by the mechanical friction of therocking beam drives. The inboard location of motor/generator 900 makesit more difficult to cool motor/generator 900, thereby increasing theeffects of heat produced by motor/generator 900 as well as heat absorbedby motor/generator 900 from the rocking beam drives. This may lead tooverheating, and ultimately failure of motor/generator 900.

Referring to both FIGS. 8 and 9, the inboard positioning ofmotor/generator 900 may also lead to an unequilateral configuration ofpistons 802, 804, 806, and 808, since pistons 802, 804, 806, and 808 arecoupled to rocking beam drives 810 and 812, respectively, and anyincrease in distance would also result in an increase in distancebetween pistons 802, 804, and pistons 806 and 808. An unequilateralarrangement of pistons may lead to inefficiencies in burner and heaterhead thermodynamic operation, which, in turn, may lead to a decrease inoverall engine efficiency. Additionally, an unequilateral arrangement ofpistons may lead to larger heater head and combustion chamberdimensions.

The exemplary embodiment of the motor/generator arrangement is shown inFIG. 10A. As shown in FIG. 10A, the motor/generator 1000 is positionedoutboard from rocking beam drives 1010 and 1012 (shown as 810 and 812 inFIG. 8) and at an end of crankshaft 1006. The outboard position allowsfor a motor/generator 1000 with a larger length and diameter squaredthan the “pancake” motor/generator described above (shown as 900 in FIG.9). As previously stated, the capacity of motor/generator 1000 isproportional to its length and diameter squared, and since outboardmotor/generator 1000 may have a larger length and diameter squared, theoutboard motor/generator 1000 configuration shown in FIG. 10A may allowfor the use of a higher capacity motor/generator in conjunction withengine.

By placing motor/generator 1000 outboard of drives 1010 and 1012 asshown in the embodiment in FIG. 10A, motor/generator 1000 is not exposedto heat generated by the mechanical friction of drives 1010 and 1012.Also, the outboard position of motor/generator 1000 makes it easier tocool the motor/generator, thereby allowing for more mechanical enginecycles per a given amount of time, which in turn allows for higheroverall engine performance.

Also, as motor/generator 1000 is positioned outside and not positionedbetween drives 1010 and 1012, rocking beam drives 1010 and 1012 may beplaced closer together thereby allowing the pistons which are coupled todrives 1010 and 1012 to be placed in an equilateral arrangement. In someembodiments, depending on the burner type used, particularly in the caseof a single burner embodiment, equilateral arrangement of pistons allowsfor higher efficiencies in burner and heater head thermodynamicoperation, which in turn allows higher overall engine performance.Equilateral arrangement of pistons also advantageously allows forsmaller heater head and combustion chamber dimensions.

Referring again to FIGS. 8 and 9, crankshaft 814 may have concentricends 902 and 904, which in one embodiment are crank journals, and invarious other embodiments, may be, but are not limited to, bearings.Each concentric end 902, 904 has a crankpin 816, 818 respectively, thatmay be offset from a crankshaft center axis. At least one counterweight906 may be placed at either end of crankshaft 814 (shown as 1006 in FIG.10A), to counterbalance any instability the crankshaft 814 mayexperience. This crankshaft configuration in combination with therocking beam drive described above allows the pistons (shown as 802,804, 806, and 808 in FIG. 8) to do work with one rotation of thecrankshaft 814. This characteristic will be further explained below. Inother embodiments, a flywheel (not shown) may be placed on crankshaft814 (shown as 1006 in FIG. 10A) to decrease fluctuations of angularvelocity for a more constant speed.

Still referring to FIGS. 8 and 9, in some embodiments, a cooler (notshown) may be also be positioned along the crankshaft 814 (shown as 1006in FIG. 10A) and rocking beam drives 810 and 812 (shown as 1010 and 1012in FIG. 10A) to cool the crankshaft 814 and rocking beam drives 810 and812. In some embodiments, the cooler may be used to cool the working gasin a cold chamber of a cylinder and may also be configured to cool therocking beam drive. Various embodiments of the cooler are discussed indetail below.

FIGS. 10A-10G depict some embodiments of various parts of the machine.As shown in this embodiment, crankshaft 1006 is coupled tomotor/generator 1000 via a motor/generator coupling assembly. Sincemotor/generator 1000 is mounted to crankcase 1008, pressurization ofcrankcase with a charge fluid may result in crankcase deformation, whichin turn may lead to misalignments between motor/generator 1000 andcrankshaft 1006 and cause crankshaft 1006 to deflect. Because rockingbeam drives 1010 and 1012 are coupled to crankshaft 1006, deflection ofcrankshaft 1006 may lead to failure of rocking beam drives 1010 and1012. Thus, in one embodiment of the machine, a motor/generator couplingassembly is used to couple the motor/generator 1000 to crankshaft 1006.The motor/generator coupling assembly accommodates differences inalignment between motor/generator 1000 and crankshaft 1006 which maycontribute to failure of rocking beam drives 1010 and 1012 duringoperation.

Still referring to FIGS. 10A-10G, in one embodiment, the motor/generatorcoupling assembly is a spline assembly that includes spline shaft 1004,sleeve rotor 1002 of motor/generator 1000, and crankshaft 1006. Splineshaft 1004 couples one end of crankshaft 1006 to sleeve rotor 1002.Sleeve rotor 1002 is attached to motor/generator 1000 by mechanicalmeans, such as press fitting, welding, threading, or the like. In oneembodiment, spline shaft 1004 includes a plurality of splines on bothends of the shaft. In other embodiments, spline shaft 1004 includes amiddle spline less portion 1014, which has a diameter smaller than theouter diameter or inner diameter of splined portions 1016 and 1018. Instill other embodiments, one end portion of the spline shaft 1016 hassplines that extend for a longer distance along the shaft than a secondend portion 1018 that also includes splines thereon.

In some embodiments, sleeve rotor 1002 includes an opening 1020 thatextends along a longitudinal axis of sleeve rotor 1002. The opening 1020is capable of receiving spline shaft 1004. In some embodiments, opening1020 includes a plurality of inner splines 1022 capable of engaging thesplines on one end of spline shaft 1004. The outer diameter 1028 ofinner splines 1022 may be larger than the outer diameter 1030 of thesplines on spline shaft 1004, such that the fit between inner splines1022 and the splines on spline shaft 1004 is loose (as shown in FIG.10E). A loose fit between inner splines 1022 and the splines on splineshaft 1004 contributes to maintain spline engagement between splineshaft 1004 and rotor sleeve 1002 during deflection of spline shaft 1004,which may be caused by crankcase pressurization. In other embodiments,longer splined portion 1016 of spline shaft 1004 may engage innersplines 1022 of rotor 1002.

Still referring to FIGS. 10A-10G, in some embodiments, crankshaft 1006has an opening 1024 on an end thereof, which is capable of receiving oneend of spline shaft 1004. Opening 1024 preferably includes a pluralityof inner splines 1026 that engage the splines on spline shaft 1004. Theouter diameter 1032 of inner splines 1026 may be larger than the outerdiameter 1034 of the splines on spline shaft 1004, such that the fitbetween inner splines 1026 and the splines on spline shaft 1004 is loose(as shown in FIG. 10F). As previously discussed, a loose fit betweeninner splines 1026 and the splines on spline shaft 1004 contributes tomaintain spline engagement between spline shaft 1004 and crankshaft 1006during deflection of spline shaft 1004, which may be caused by crankcasepressurization. The loose fit between the inner splines 1026 and 1022 onthe crankshaft 1006 and the sleeve rotor 1002 and the splines on thespline shaft 1004 may contribute to maintain deflection of spline shaft1004. This may allow misalignments between crankshaft 1006 and sleeverotor 1002. In some embodiments, shorter splined portion 1018 of splineshaft 1004 may engage opening 1024 of crankshaft 1006 thus preventingthese potential misalignments.

In some embodiments, opening 1020 of sleeve rotor 1002 includes aplurality of inner splines that extend the length of opening 1020. Thisarrangement contributes to spline shaft 1004 being properly insertedinto opening 1020 during assembly. This contributes to proper alignmentbetween the splines on spline shaft 1004 and the inner splines on sleeverotor 1002 being maintained.

Referring now to FIG. 4, one embodiment of the engine is shown. Here thepistons 202 and 204 of engine 300 operate between a hot chamber 404 anda cold chamber 406 of cylinders 206 and 208 respectively. Between thetwo chambers there may be a regenerator 408. The regenerator 408 mayhave variable density, variable area, and, in some embodiments, is madeof wire. The varying density and area of the regenerator may be adjustedsuch that the working gas has substantially uniform flow across theregenerator 408. Various embodiments of the regenerator 408 arediscussed in detail below, and in U.S. Pat. No. 6,591,609, issued Jul.17, 2003, to Kamen et al., and U.S. Pat. No. 6,862,883, issued Mar. 8,2005, to Kamen et al., which are herein incorporated by reference intheir entireties. When the working gas passes through the hot chamber404, a heater head 410 may heat the gas causing the gas to expand andpush pistons 202 and 204 towards the cold chamber 406, where the gascompresses. As the gas compresses in the cold chamber 406, pistons 202and 204 may be guided back to the hot chamber to undergo the Stirlingcycle again. The heater head 410 may be a pin head (as shown in FIGS.52A through 53B), a fin head (as shown in FIGS. 56A through 56C), afolded fin head (as shown in FIGS. 56A through 56C), heater tubes asshown in FIG. 4 (also shown as 2904 in FIG. 29), or any other heaterhead embodiment known, including, but not limited to, those describedbelow. Various embodiments of heater head 410 are discussed in detailbelow, and in U.S. Pat. No. 6,381,958, issued May 7, 2002, to Kamen etal., U.S. Pat. No. 6,543,215, issued Apr. 8, 2003, to Langenfeld et al.,U.S. Pat. No. 6,966,182, issued Nov. 22, 2005, to Kamen et al, and U.S.Pat. No. 7,308,787, issued Dec. 18, 2007, to LaRocque et al., which areherein incorporated by reference in their entireties.

In some embodiments, a cooler 412 may be positioned alongside cylinders206 and 208 to further cool the gas passing through to the cold chamber406. Various embodiments of cooler 412 are discussed in detail in theproceeding sections, and in U.S. Pat. No. 7,325,399, issued Feb. 5,2008, to Strimling et al, which is herein incorporated by reference inits entirety.

In some embodiments, at least one piston seal 414 may be positioned onpistons 202 and 204 to seal the hot section 404 off from the coldsection 406. Additionally, at least one piston guide ring 416 may bepositioned on pistons 202 and 204 to help guide the pistons' motion intheir respective cylinders. Various embodiments of piston seal 414 andguide ring 416 are described in detail below, and in U.S. patentapplication Ser. No. 10/175,502, filed Jun. 19, 2002, published Feb. 6,2003 (now abandoned), which is herein incorporated by reference in itsentirety.

In some embodiments, at least one piston rod seal 418 may be placedagainst piston rods 224 and 228 to prevent working gas from escapinginto the crankcase 400, or alternatively into airlock space 420. Thepiston rod seal 418 may be an elastomer seal, or a spring-loaded seal.Various embodiments of the piston rod seal 418 are discussed in detailbelow.

In some embodiments, the airlock space may be eliminated, for example,in the rolling diaphragm and/or bellows embodiments described in moredetail below. In those cases, the piston rod seals 224 and 228 seal theworking space from the crankcase.

In some embodiments, at least one rolling diaphragm/bellows 422 may belocated along piston rods 224 and 228 to prevent airlock gas fromescaping into the crankcase 400. Various embodiments of rollingdiaphragm 422 are discussed in more detail below.

Although FIG. 4 shows a cross section of engine 300 depicting only twopistons and one rocking beam drive, it is to be understood that theprinciples of operation described herein may apply to a four cylinder,double rocking beam drive engine, as designated generally by numeral 800in FIG. 8.

Piston Operation

Referring now to FIGS. 8 and 11, FIG. 11 shows the operation of pistons802, 804, 806, and 808 during one revolution of crankshaft 814. With a ¼revolution of crankshaft 814, piston 802 is at the top of its cylinder,otherwise known as top dead center, piston 806 is in upward mid-stroke,piston 804 is at the bottom of its cylinder, otherwise known as bottomdead center, and piston 808 is in downward mid-stroke. With a ½revolution of crankshaft 814, piston 802 is in downward mid-stroke,piston 806 is at top dead center, piston 804 is in upward mid-stroke,and piston 808 is at bottom dead center. With ¾ revolution of crankshaft814, piston 802 is at bottom dead center, piston 806 is in downwardmid-stroke, piston 804 is at top dead center, and piston 808 is inupward midstroke. Finally, with a full revolution of crankshaft 814,piston 802 is in upward midstroke, piston 806 is at bottom dead center,piston 804 is in downward mid-stroke, and piston 808 is at top deadcenter. During each ¼ revolution, there is a 90 degree phase differencebetween pistons 802 and 806, a 180 degree phase difference betweenpistons 802 and 804, and a 270 degree phase difference between pistons802 and 808. FIG. 12A illustrates the relationship of the pistons beingapproximately 90 degrees out of phase with the preceding and succeedingpiston. Additionally, FIG. 11 shows the exemplary embodiment machinemeans of transferring work. Thus, work is transferred from piston 802 topiston 806 to piston 804 to piston 808 so that with a full revolution ofcrankshaft 814, all pistons have exerted work by moving from the top tothe bottom of their respective cylinders.

Referring now to FIG. 11, together with FIGS. 12A-12C, illustrate the 90degree phase difference between the pistons in the exemplary embodiment.Referring now to FIG. 12A, although the cylinders are shown in a linearpath, this is for illustration purposes only. In the exemplaryembodiment of a four cylinder Stirling cycle machine, the flow path ofthe working gas contained within the cylinder working space follows afigure eight pattern. Thus, the working spaces of cylinders 1200, 1202,1204, and 1206 are connected in a figure eight pattern, for example,from cylinder 1200 to cylinder 1202 to cylinder 1204 to cylinder 1208,the fluid flow pattern follows a figure eight. Still referring to FIG.12A, an unwrapped view of cylinders 1200, 1202, 1204, and 1206, takenalong the line B-B (shown in FIG. 12C) is illustrated. The 90 degreephase difference between pistons as described above allows for theworking gas in the warm section 1212 of cylinder 1204 to be delivered tothe cold section 1222 of cylinder 1206. As piston 802 and 808 are 90degrees out of phase, the working gas in the warm section 1214 ofcylinder 1206 is delivered to the cold section 1216 of cylinder 1200. Aspiston 802 and piston 806 are also 90 degrees out of phase, the workinggas in the warm section 1208 of cylinder 1200 is delivered to the coldsection 1218 of cylinder 1202. And as piston 804 and piston 806 are also90 degrees out of phase, so the working gas in the warm section 1210 ofcylinder 1202 is delivered to the cold section 1220 of cylinder 1204.Once the working gas of a warm section of a first cylinder enters thecold section of a second cylinder, the working gas begins to compress,and the piston within the second cylinder, in its down stroke,thereafter forces the compressed working gas back through a regenerator1224 and heater head 1226 (shown in FIG. 12B), and back into the warmsection of the first cylinder. Once inside the warm section of the firstcylinder, the gas expands and drives the piston within that cylinderdownward, thus causing the working gas within the cold section of thatfirst cylinder to be driven through the preceding regenerator and heaterhead, and into the cylinder. This cyclic transmigration characteristicof working gas between cylinders 1200, 1202, 1204, and 1206 is possiblebecause pistons 802, 804, 806, and 808 are connected, via drives 810 and812, to a common crankshaft 814 (shown in FIG. 11), in such a way thatthe cyclical movement of each piston is approximately 90 degrees inadvance of the movement of the proceeding piston, as depicted in FIG.12A.

Rolling Diaphragm, Metal Bellows, Airlock, and Pressure Regulator

In some embodiments of the Stirling cycle machine, lubricating fluid isused. To prevent the lubricating fluid from escaping the crankcase, aseal is used.

Referring now to FIGS. 13A-15, some embodiments of the Stirling cyclemachine include a fluid lubricated rocking beam drive that utilizes arolling diaphragm 1300 positioned along the piston rod 1302 to preventlubricating fluid from escaping the crankcase, not shown, but thecomponents that are housed in the crankcase are represented as 1304, andentering areas of the engine that may be damaged by the lubricatingfluid. It is beneficial to contain the lubricating fluid for iflubricating fluid enters the working space, not shown, but thecomponents that are housed in the working space are represented as 1306,it would contaminate the working fluid, come into contact with theregenerator 1308, and may clog the regenerator 1308. The rollingdiaphragm 1300 may be made of an elastomer material, such as rubber orrubber reinforced with woven fabric or non-woven fabric to providerigidity. The rolling diaphragm 1300 may alternatively be made of othermaterials, such as fluorosilicone or nitrile with woven fabric ornon-woven fabric. The rolling diaphragm 1300 may also be made of carbonnanotubes or chopped fabric, which is non-woven fabric with fibers ofpolyester or KEVLAR®, for example, dispersed in an elastomer. In thesome embodiments, the rolling diaphragm 1300 is supported by the topseal piston 1328 and the bottom seal piston 1310. In other embodiments,the rolling diaphragm 1300 as shown in FIG. 13A is supported via notchesin the top seal piston 1328.

In some embodiments, a pressure differential is placed across therolling diaphragm 1300 such that the pressure above the seal 1300 isdifferent from the pressure in the crankcase 1304. This pressuredifferential inflates seal 1300 and allows seal 1300 to act as a dynamicseal as the pressure differential ensures that rolling diaphragmmaintains its form throughout operation. FIG. 13A, and FIGS. 13C-13Hillustrate how the pressure differential effects the rolling diaphragm.The pressure differential causes the rolling diaphragm 1300 to conformto the shape of the bottom seal piston 1310 as it moves with the pistonrod 1302, and prevents separation of the seal 1300 from a surface of thepiston 1310 during operation. Such separation may cause seal failure.The pressure differential causes the rolling diaphragm 1300 to maintainconstant contact with the bottom seal piston 1310 as it moves with thepiston rod 1302. This occurs because one side of the seal 1300 willalways have pressure exerted on it thereby inflating the seal 1300 toconform to the surface of the bottom seal piston 1310. In someembodiments, the top seal piston 1328 ‘rolls over’ the corners of therolling diaphragm 1300 that are in contact with the bottom seal piston1310, so as to further maintain the seal 1300 in contact with the bottomseal piston 1310. In the exemplary embodiment, the pressure differentialis in the range of 10 to 15 PSI. The smaller pressure in the pressuredifferential is preferably in crankcase 1304, so that the rollingdiaphragm 1300 may be inflated into the crankcase 1304. However, inother embodiments, the pressure differential may have a greater orsmaller range of value.

The pressure differential may be created by various methods including,but not limited to, the use of the following: a pressurized lubricationsystem, a pneumatic pump, sensors, an electric pump, by oscillating therocking beam to create a pressure rise in the crankcase 1304, bycreating an electrostatic charge on the rolling diaphragm 1300, or othersimilar methods. In some embodiments, the pressure differential iscreated by pressurizing the crankcase 1304 to a pressure that is belowthe mean pressure of the working space 1306. In some embodiments thecrankcase 1304 is pressurized to a pressure in the range of 10 to 15 PSIbelow the mean pressure of the working space 1306, however, in variousother embodiments, the pressure differential may be smaller or greater.Further detail regarding the rolling diaphragm is included below.

Referring now to FIGS. 13C, 13G, and 13H, however, another embodiment ofthe Stirling machine is shown, wherein airlock space 1312 is locatedbetween working space 1306 and crankcase 1304. Airlock space 1312maintains a constant volume and pressure necessary to create thepressure differential necessary for the function of rolling diaphragm1300 as described above. In one embodiment, airlock 1312 is notabsolutely sealed off from working space 1306, so the pressure ofairlock 1312 is equal to the mean pressure of working space 1306. Thus,in some embodiments, the lack of an effective seal between the workingspace and the crankcase contributes to the need for an airlock space.Thus, the airlock space, in some embodiments, may be eliminated by amore efficient and effective seal.

During operation, the working space 1306 mean pressure may vary so as tocause airlock 1312 mean pressure to vary as well. One reason thepressure may tend to vary is that during operation the working space mayget hotter, which in turn may increase the pressure in the workingspace, and consequently in the airlock as well since the airlock andworking space are in fluid communication. In such a case, the pressuredifferential between airlock 1312 and crankcase 1304 will also vary,thereby causing unnecessary stresses in rolling diaphragms 1300 that maylead to seal failure. Therefore, some embodiments of the machine, themean pressure within airlock 1312 is regulated so as to maintain aconstant desired pressure differential between airlock 1312 andcrankcase 1304, and ensuring that rolling diaphragms 1300 stay inflatedand maintains their form. In some embodiments, a pressure transducer isused to monitor and manage the pressure differential between the airlockand the crankcase, and regulate the pressure accordingly so as tomaintain a constant pressure differential between the airlock and thecrankcase. Various embodiments of the pressure regulator that may beused are described in further detail below, and in U.S. Pat. No.7,310,945, issued Dec. 25, 2007, to Gurski et al., which is hereinincorporated by reference in its entirety.

A constant pressure differential between the airlock 1312 and crankcase1304 may be achieved by adding or removing working fluid from airlock1312 via a pump or a release valve. Alternatively, a constant pressuredifferential between airlock 1312 and crankcase 1304 may be achieved byadding or removing working fluid from crankcase 1304 via a pump or arelease valve. The pump and release valve may be controlled by thepressure regulator. Working fluid may be added to airlock 1312 (orcrankcase 1304) from a separate source, such as a working fluidcontainer, or may be transferred over from crankcase 1304. Shouldworking fluid be transferred from crankcase 1304 to airlock 1312, it maybe desirable to filter the working fluid before passing it into airlock1312 so as to prevent any lubricant from passing from crankcase 1304into airlock 1312, and ultimately into working space 1306, as this mayresult in engine failure.

In some embodiments of the machine, crankcase 1304 may be charged with afluid having different thermal properties than the working fluid. Forexample, where the working gas is helium or hydrogen, the crankcase maybe charged with argon. Thus, the crankcase is pressurized. In someembodiments, helium is used, but in other embodiments, any inert gas, asdescribed herein, may be used. Thus, the crankcase is a wet pressurizedcrankcase in the exemplary embodiment. In other embodiments where alubricating fluid is not used, the crankcase is not wet.

In the exemplary embodiments, rolling diaphragms 1300 do not allow gasor liquid to pass through them, which allows working space 1306 toremain dry and crankcase 1304 to be wet sumped with a lubricating fluid.Allowing a wet sump crankcase 1304 increases the efficiency and life ofthe engine as there is less friction in rocking beam drives 1316. Insome embodiments, the use of roller bearings or ball bearings in drives1316 may also be eliminated with the use of lubricating fluid androlling diaphragms 1300. This may further reduce engine noise andincrease engine life and efficiency.

FIGS. 14A-14E show cross sections of various embodiments of the rollingdiaphragm (shown as 1400, 1410, 1412, 1422 and 1424) configured to bemounted between top seal piston and bottom seal piston (shown as 1328and 1310 in FIGS. 13A and 13H), and between a top mounting surface and abottom mounting surface (shown as 1320 and 1318 in FIG. 13A). In someembodiments, the top mounting surface may be the surface of an airlockor working space, and the bottom mounting surface may be the surface ofa crankcase.

FIG. 14A shows one embodiment of the rolling diaphragm 1400, where therolling diaphragm 1400 includes a flat inner end 1402 that may bepositioned between a top seal piston and a bottom seal piston, so as toform a seal between the top seal piston and the bottom seal piston. Therolling diaphragm 1400 also includes a flat outer end 1404 that may bepositioned between a top mounting surface and a bottom mounting surface,so as to form a seal between the top mounting surface and the bottommounting surface. FIG. 14B shows another embodiment of the rollingdiaphragm, wherein rolling diaphragm 1410 may include a plurality ofbends 1408 leading up to flat inner end 1406 to provide for additionalsupport and sealing contact between the top seal piston and the bottomseal piston. FIG. 14C shows another embodiment of the rolling diaphragm,wherein rolling diaphragm 1412 includes a plurality of bends 1416leading up to flat outer end 1414 to provide for additional support andsealing contact between the top mounting surface and the bottom mountingsurface.

FIG. 14D shows another embodiment of the rolling diaphragm where rollingdiaphragm 1422 includes a bead along an inner end 1420 thereof, so as toform an ‘o-ring’ type seal between a top seal piston and a bottom sealpiston, and a bead along an outer end 1418 thereof, so as to form an‘o-ring’ type seal between a bottom mounting surface and a top mountingsurface. FIG. 14E shows another embodiment of the rolling diaphragm,wherein rolling diaphragm 1424 includes a plurality of bends 1428leading up to beaded inner end 1426 to provide for additional supportand sealing contact between the top seal piston and the bottom sealpiston. Rolling diaphragm 1424 may also include a plurality of bends1430 leading up to beaded outer end 1432 to provide for additionalsupport and sealing contact between the top seal piston and the bottomseal piston.

Although FIGS. 14A through 14E depict various embodiments of the rollingdiaphragm, it is to be understood that rolling diaphragms may be held inplace by any other mechanical means known in the art.

Referring now to FIG. 15A, a cross section shows one embodiment of therolling diaphragm embodiment. A metal bellows 1500 is positioned along apiston rod 1502 to seal off a crankcase (shown as 1304 in FIG. 13G) froma working space or airlock (shown as 1306 and 1312 in FIG. 13G). Metalbellows 1500 may be attached to a top seal piston 1504 and a stationarymounting surface 1506. Alternatively, metal bellows 1500 may be attachedto a bottom seal piston (not shown), and a top stationary mountingsurface. In one embodiment the bottom stationary mounting surface may bea crankcase surface or an inner airlock or working space surface, andthe top stationary mounting surface may be an inner crankcase surface,or an outer airlock or working space surface. Metal bellows 1500 may beattached by welding, brazing, or any mechanical means known in the art.

FIGS. 15B-15G depict a perspective cross sectional view of variousembodiments of the metal bellows, wherein the metal bellows is a weldedmetal bellows 1508. In some embodiments of the metal bellows, the metalbellows is preferably a micro-welded metal bellows. In some embodiments,the welded metal bellows 1508 includes a plurality of diaphragms 1510,which are welded to each other at either an inner end 1512 or an outerend 1514, as shown in FIGS. 15C and 15D. In some embodiments, diaphragms1510 may be crescent shaped 1516, flat 1518, rippled 1520, or any othershape known in the art.

Additionally, the metal bellows may alternatively be formed mechanicallyby means such as die forming, hydroforming, explosive hydroforming,hydramolding, or any other means known in the art.

The metal bellows may be made of any type of metal, including but notlimited to, steel, stainless steel, stainless steel 374, AM-350stainless steel, Inconel, Hastelloy, Haynes, titanium, or any otherhigh-strength, corrosion-resistant material.

In one embodiment, the metal bellows used are those available fromSenior Aerospace Metal Bellows Division, Sharon, Mass., or American BOA,Inc., Cumming, Ga.

Rolling Diaphragm and/or Bellows Embodiments

Various embodiments of the rolling diaphragm and/or bellows, whichfunction to seal, are described above. Further embodiments will beapparent to those of skill in the art based on the description above andthe additional description below relating to the parameters of therolling diaphragm and/or bellows.

In some embodiments, the pressure atop the rolling diaphragm or bellows,in the airlock space or airlock area (both terms are usedinterchangeably), is the mean-working-gas pressure for the machine,which, in some embodiments is an engine, while the pressure below therolling diaphragm and/or bellows, in the crankcase area, isambient/atmospheric pressure. In these embodiments, the rollingdiaphragm and/or bellows is required to operate with as much as 3000 psiacross it (and in some embodiments, up to 1500 psi or higher). In thiscase, the rolling diaphragm and/or bellows seal forms the working gas(helium, hydrogen, or otherwise) containment barrier for the machine(engine in the exemplary embodiment). Also, in these embodiments, theneed for a heavy, pressure-rated, structural vessel to contain thebottom end of the engine is eliminated, since it is now required tosimply contain lubricating fluid (oil is used as a lubricating fluid inthe exemplary embodiment) and air at ambient pressure, like aconventional internal combustion (“IC”) engine.

The capability to use a rolling diaphragm and/or bellows seal with suchan extreme pressure across it depends on the interaction of severalparameters. Referring now to FIG. 15H, an illustration of the actualload on the rolling diaphragm or bellows material is shown. As shown,the load is a function of the pressure differential and the annular gaparea for the installed rolling diaphragm or bellows seal.

Region 1 represents the portions of the rolling diaphragm and/or bellowsthat are in contact with the walls formed by the piston and cylinder.The load is essentially a tensile load in the axial direction, due tothe pressure differential across the rolling diaphragm and/or bellows.This tensile load due to the pressure across the rolling diaphragmand/or bellows can be expressed as:L _(t) =P _(d) *A _(a)

Where

-   -   L_(t)=Tensile Load and    -   P_(d)=Pressure Differential    -   A_(a)=Annular Area

andA _(a) =p/4*(D ² −d ²)

Where

-   -   D=Cylinder Bore and    -   d=Piston Diameter

The tensile component of stress in the bellows material can beapproximated as:S _(t) =L _(t)/(p*(D+d)*t _(b))Which reduces to:S _(t) =Pa/4*(D−d)/tb

Later, we will show the relationship of radius of convolution, R_(c), toCylinder bore (D) and Piston Diameter (d) to be defined as:R _(c)=(D−d)/4

So, this formula for St reduces to its final form:S _(t) =P _(d) *R _(c) /t _(b)

Where

-   -   t_(b)=thickness of bellows material

Still referring to FIG. 15H, Region 2 represents the convolution. As therolling diaphragm and/or bellows material turns the corner, in theconvolution, the hoop stress imposed on the rolling diaphragm and/orbellows material may be calculated. For the section of the bellowsforming the convolution, the hoop component of stress can be closelyapproximated as:S _(h) =P _(d) *R _(c) /t _(b)

The annular gap that the rolling diaphragm and/or bellows rolls withinis generally referred to as the convolution area. The rolling diaphragmand/or bellows fatigue life is generally limited by the combined stressfrom both the tensile (and hoop) load, due to pressure differential, aswell as the fatigue due to the bending as the fabric rolls through theconvolution. The radius that the fabric takes on during this ‘rolling’is defined here as the radius of convolution, Rc.R _(c)=(D−d)/4

The bending stress, Sb, in the rolling diaphragm and/or bellows materialas it rolls through the radius of convolution, Rc, is a function of thatradius, as well as the thickness of the materials in bending. For afiber-reinforced material, the stress in the fibers themselves (duringthe prescribed deflection in the exemplary embodiments) is reduced asthe fiber diameter decreases. The lower resultant stress for the samelevel of bending allows for an increased fatigue life limit. As thefiber diameter is further reduced, flexibility to decrease the radius ofconvolution Rc is achieved, while keeping the bending stress in thefiber under its endurance limit. At the same time, as Rc decreases, thetensile load on the fabric is reduced since there is less unsupportedarea in the annulus between the piston and cylinder. The smaller thefiber diameter, the smaller the minimum Rc, the smaller the annulararea, which results in a higher allowable pressure differential.

For bending around a prescribed radius, the bending moment isapproximated by:M=E*I/R

Where:

-   -   M=Bending Moment    -   E=Elastic Modulus    -   I=Moment of Inertia    -   R=Radius of Bend

Classical bending stress, S_(b), is calculated as:S _(b) =M*Y/I

Where:

-   -   Y=Distance above neutral axis of bending

Substituting yields:S _(b)=(E*I/R)*Y/IS _(b) =E*Y/R

Assuming bending is about a central neutral axis:Y _(max) =t _(b)/2S _(b) =E*t _(b)/(2*R)

In some embodiments, rolling diaphragm and/or bellows designs for highcycle life are based on geometry where the bending stress imposed iskept about one order of magnitude less than the pressure-based loading(hoop and axial stresses). Based on the equation: Sb=E*tb/(2*R), it isclear that minimizing tb in direct proportion to Rc should not increasethe bending stress. The minimum thickness for the exemplary embodimentsof the rolling diaphragm and/or bellows material or membrane is directlyrelated to the minimum fiber diameter that is used in the reinforcementof the elastomer. The smaller the fibers used, the smaller resultant Rcfor a given stress level.

Another limiting component of load on the rolling diaphragm and/orbellows is the hoop stress in the convolution (which is theoreticallythe same in magnitude as the axial load while supported by the piston orcylinder). The governing equation for that load is as follows:Sh=P _(d) *Rc/tb

Thus, if Rc is decreased in direct proportion to tb, then there is noincrease of stress on the membrane in this region. However, if thisratio is reduced in a manner that decreases Rc to a greater ratio thantb then parameters must be balanced. Thus, decreasing tb with respect toRc requires the rolling diaphragm and/or bellows to carry a heavierstress due to pressure, but makes for a reduced stress level due tobending. The pressure-based load is essentially constant, so this may befavorable—since the bending load is cyclic, therefore it is the bendingload component that ultimately limits fatigue life.

For bending stress reduction, tb ideally should be at a minimum, and Rcideally should be at a maximum. E ideally is also at a minimum. For hoopstress reduction, Rc ideally is small, and tb ideally is large.

Thus, the critical parameters for the rolling diaphragm and/or bellowsmembrane material are:

E, Elastic Modulus of the membrane material;

tb, membrane thickness (and/or fiber diameter);

Sut, Ultimate tensile strength of the rolling diaphragm and/or bellows;and

Slcf, The limiting fatigue strength of the rolling diaphragm and/orbellows.

Thus, from E, tb and Sut, the minimum acceptable Rc may be calculated.Next, using Rc, Slcf, and tb, the maximum Pd may be calculates. Rc maybe adjusted to shift the bias of load (stress) components between thesteady state pressure stress and the cyclic bending stress. Thus, theideal rolling diaphragm and/or bellows material is extremely thin,extremely strong in tension, and very limber in flexion.

Thus, in some embodiments, the rolling diaphragm and/or bellows material(sometimes referred to as a “membrane”), is made from carbon fibernanotubes. However, additional small fiber materials may also be used,including, but not limited to nanotube fibers that have been braided,nanotube untwisted yarn fibers, or any other conventional materials,including but not limited to KEVLAR, glass, polyester, synthetic fibersand any other material or fiber having a desirable diameter and/or otherdesired parameters as described in detail above.

Piston Seals and Piston Rod Seals

Referring now to FIG. 13G, an embodiment of the machine is shown whereinan engine 1326, such as a Stirling cycle engine, includes at least onepiston rod seal 1314, a piston seal 1324, and a piston guide ring 1322,(shown as 1616 in FIG. 16). Various embodiments of the piston seal 1324and the piston guide ring 1322 are further discussed below, and in U.S.patent application Ser. No. 10/175,502 (now abandoned), which, asmentioned before, is incorporated by reference.

FIG. 16 shows a partial cross section of the piston 1600, driven alongthe central axis 1602 of cylinder, or the cylinder 1604. The piston seal(shown as 1324 in FIG. 13G) may include a seal ring 1606, which providesa seal against the contact surface 1608 of the cylinder 1604. Thecontact surface 1608 is typically a hardened metal (preferably 58-62 RC)with a surface finish of 12 RMS or smoother. The contact surface 1608may be metal which has been case hardened, such as 8260 hardened steel,which may be easily case hardened and may be ground and/or honed toachieve a desired finish. The piston seal may also include a backingring 1610, which is sprung to provide a thrust force against the sealring 1606 thereby providing sufficient contact pressure to ensuresealing around the entire outward surface of the seal ring 1606. Theseal ring 1606 and the backing ring 1610 may together be referred to asa piston seal composite ring. In some embodiments, the at least onepiston seal may seal off a warm portion of cylinder 1604 from a coldportion of cylinder 1604.

Referring now to FIG. 17, some embodiments include a piston rod seal(shown as 1314 in FIG. 13G) mounted in the piston rod cylinder wall1700, which, in some embodiments, may include a seal ring 1706, whichprovides a seal against the contact surface 1708 of the piston rod 1604(shown as 1302 in FIG. 13G). The contact surface 1708 in someembodiments is a hardened metal (preferably 58-62 RC) with a surfacefinish of 12 RMS or smoother. The contact surface 1708 may be metalwhich has been case hardened, such as 8260 hardened steel, which may beeasily case hardened and may be ground and/or honed to achieve a desiredfinish. The piston seal may also include a backing ring 1710, which issprung to provide a radial or hoop force against the seal ring 1706thereby providing sufficient contact hoop stress to ensure sealingaround the entire inward surface of seal ring 1706. The seal ring 1706and the backing ring 1710 may together be referred to as a piston rodseal composite ring.

In some embodiments, the seal ring and the backing ring may bepositioned on a piston rod, with the backing exerting an outwardpressure on the seal ring, and the seal ring may come into contact witha piston rod cylinder wall 1702. These embodiments require a largerpiston rod cylinder length than the previous embodiment. This is becausethe contact surface on the piston rod cylinder wall 1702 will be longerthan in the previous embodiment, where the contact surface 1708 lies onthe piston rod itself. In yet another embodiment, piston rod seals maybe any functional seal known in the art including, but not limited to,an o-ring, a graphite clearance seal, graphite piston in a glasscylinder, or any air pot, or a spring energized lip seal. In someembodiments, anything having a close clearance may be used, in otherembodiments, anything having interference, for example, a seal, is used.In the exemplary embodiment, a spring energized lip seal is used. Anyspring energized lip seal may be used, including those made by BAL SEALEngineering, Inc., Foothill Ranch, Calif. In some embodiments, the sealused is a BAL SEAL Part Number X558604.

The material of the seal rings 1606 and 1706 is chosen by considering abalance between the coefficient of friction of the seal rings 1606 and1706 against the contact surfaces 1608 and 1708, respectively, and thewear on the seal rings 1606 and 1706 it engenders. In applications inwhich piston lubrication is not possible, such as at the high operatingtemperatures of a Stirling cycle engine, the use of engineering plasticrings is used. The embodiments of the composition include a nylon matrixloaded with a lubricating and wear-resistant material. Examples of suchlubricating materials include PTFE/silicone, PTFE, graphite, etc.Examples of wear-resistant materials include glass fibers and carbonfibers. Examples of such engineering plastics are manufactured by LNPEngineering Plastics, Inc. of Exton, Pa. Backing rings 1610 and 1710 ispreferably metal.

The fit between the seal rings 1606 and 1706 and the seal ring grooves1612 and 1712, respectively, is preferably a clearance fit (about0.002″), while the fit of the backing rings 1610 and 1710 is preferablya looser fit, of the order of about 0.005″ in some embodiments. The sealrings 1606 and 1706 provide a pressure seal against the contact surfaces1608 and 1708, respectively, and also one of the the surfaces 1614 and1714 of the seal ring grooves 1612 and 1712, respectively, depending onthe direction of the pressure difference across the rings 1606 and 1706and the direction of the piston 1600 or the piston rod 1704 travel.

FIGS. 18A and 18B show that if the backing ring 1820 is essentiallycircularly symmetrical, but for the gap 1800, it will assume, uponcompression, an oval shape, as shown by the dashed backing ring 1802.The result may be an uneven radial or hoop force (depicted by arrows1804) exerted on the seal ring (not shown, shown as 1606 and 1706 inFIGS. 16 and 17), and thus an uneven pressure of the seal rings againstthe contact surfaces (not shown, shown as 1608 and 1708 in FIGS. 16 and17) respectively, causing uneven wear of the seal rings and in somecases, failure of the seals.

A solution to the problem of uneven radial or hoop force exerted by thepiston seal backing ring 1820, in accordance with an embodiment, is abacking ring 1822 having a cross-section varying with circumferentialdisplacement from the gap 1800, as shown in FIGS. 18C and 18D. Atapering of the width of the backing ring 1822 is shown from theposition denoted by numeral 1806 to the position denoted by numeral1808. Also shown in FIGS. 18C and 18D is a lap joint 1810 providing forcircumferential closure of the seal ring 1606. As some seals will wearsignificantly over their lifetime, the backing ring 1822 should providean even pressure (depicted by numeral 1904 in FIG. 19B) of a range ofmovement. The tapered backing ring 1822 shown in FIGS. 18C and 18D mayprovide this advantage.

FIGS. 19A and 19B illustrate another solution to the problem of unevenradial or hoop force of the piston seal ring against the pistoncylinder, in accordance with some embodiments. As shown in FIG. 19A,backing ring 1910 is fashioned in an oval shape, so that uponcompression within the cylinder, the ring assumes the circular shapeshown by dashed backing ring 1902. A constant contact pressure betweenthe seal ring and the cylinder contact surface may thus be provided byan even radial force 1904 of backing ring 1902, as shown in FIG. 19B.

A solution to the problem of uneven radial or hoop force exerted by thepiston rod seal backing ring, in accordance with some embodiments, is abacking ring 1824 having a cross-section varying with circumferentialdisplacement from gap 1812, as shown in FIGS. 18E and 18F. A tapering ofthe width of backing ring 1824 is shown from the position denoted bynumeral 1814 to the position denoted by numeral 1816. Also shown inFIGS. 18E and 18F is a lap joint 1818 providing for circumferentialclosure of seal ring 1706. As some seals will wear significantly overtheir lifetime, backing ring 1824 should provide an even pressure(depicted by numeral 2004 in FIG. 20B) of a range of movement. Thetapered backing ring 1824 shown in FIGS. 18E and 18F may provide thisadvantage.

FIGS. 20A and 20B illustrate another solution to the problem of unevenradial or hoop force of the piston rod seal ring against the piston rodcontact surface, in accordance with some embodiments. As shown in FIG.20A, backing ring (shown by dashed backing ring 2000) is fashioned as anoval shape, so that upon expansion within the cylinder, the ring assumesthe circular shape shown by backing ring 2002. A constant contactpressure between the seal ring 1706 and the cylinder contact surface maythus be provided by an even radial thrust force 2004 of backing ring2002, as shown in FIG. 20B.

Referring again to FIG. 16, at least one guide ring 1616 may also beprovided, in accordance with some embodiments, for bearing any side loadon piston 1600 as it moves up and down the cylinder 1604. Guide ring1616 is also preferably fabricated from an engineering plastic materialloaded with a lubricating material. A perspective view of guide ring1616 is shown in FIG. 21. An overlapping joint 2100 is shown and may bediagonal to the central axis of guide ring 1616.

Lubricating Fluid Pump and Lubricating Fluid Passageways

Referring now to FIG. 22, a representative illustration of oneembodiment of the engine 2200 for the machine is shown having a rockingbeam drive 2202 and lubricating fluid 2204. In some embodiments, thelubricating fluid is oil. The lubricating fluid is used to lubricateengine parts in the crankcase 2206, such as hydrodynamic pressure fedlubricated bearings. Lubricating the moving parts of the engine 2200serves to further reduce friction between engine parts and furtherincrease engine efficiency and engine life. In some embodiments,lubricating fluid may be placed at the bottom of the engine, also knownas an oil sump, and distributed throughout the crankcase. Thelubricating fluid may be distributed to the different parts of theengine 2200 by way of a lubricating fluid pump, wherein the lubricatingfluid pump may collect lubricating fluid from the sump via a filteredinlet. In the exemplary embodiment, the lubricating fluid is oil andthus, the lubricating fluid pump is herein referred to as an oil pump.However, the term “oil pump” is used only to describe the exemplaryembodiment and other embodiments where oil is used as a lubricatingfluid, and the term shall not be construed to limit the lubricatingfluid or the lubricating fluid pump.

Referring now to FIGS. 23A and 23B, one embodiment of the engine isshown, wherein lubricating fluid is distributed to different parts ofthe engine 2200 that are located in the crankcase 2206 by a mechanicaloil pump 2208. The oil pump 2208 may include a drive gear 2210 and anidle gear 2212. In some embodiments, the mechanical oil pump 2208 may bedriven by a pump drive assembly. The pump drive assembly may include adrive shaft 2214 coupled to a drive gear 2210, wherein the drive shaft2214 includes an intermediate gear 2216 thereon. The intermediate gear2216 is preferably driven by a crankshaft gear 2220, wherein thecrankshaft gear 2220 is coupled to the primary crankshaft 2218 of theengine 2200, as shown in FIG. 24. In this configuration, the crankshaft2218 indirectly drives the mechanical oil pump 2208 via the crankshaftgear 2220, which drives the intermediate gear 2216 on the drive shaft2214, which, in turn, drives the drive gear 2210 of the oil pump 2208.

The crankshaft gear 2220 may be positioned between the crankpins 2222and 2224 of crankshaft 2218 in some embodiments, as shown in FIG. 24. Inother embodiments, the crankshaft gear 2220 may be placed at an end ofthe crankshaft 2218, as shown in FIGS. 25A-25C.

For ease of manufacturing, the crankshaft 2218 may be composed of aplurality of pieces. In these embodiments, the crankshaft gear 2220 maybe to be inserted between the crankshaft pieces during assembly of thecrankshaft.

The drive shaft 2214, in some embodiments, may be positionedperpendicularly to the crankshaft 2218, as shown in FIGS. 23A and 25A.However, in some embodiments, the drive shaft 2214 may be positionedparallel to the crankshaft 2218, as shown in FIGS. 25B and 25C.

In some embodiments, the crankshaft gear 2234 and the intermediate gear2232 may be sprockets, wherein the crankshaft gear 2234 and theintermediate gear 2232 are coupled by a chain 2226, as shown in FIGS.25C and 26C. In such an embodiments, the chain 2226 is used to drive achain drive pump (shown as 2600 in FIGS. 26A through 26C).

In some embodiments, the gear ratio between the crankshaft 2218 and thedrive shaft 2214 remains constant throughout operation. In such anembodiment, it is important to have an appropriate gear ratio betweenthe crankshaft and the drive shaft, such that the gear ratio balancesthe pump speed and the speed of the engine. This achieves a specifiedflow of lubricant required by a particular engine RPM (revolutions perminute) operating range.

In some embodiments, lubricating fluid is distributed to different partsof an engine by an electric pump. The electric pump eliminates the needfor a pump drive assembly, which is otherwise required by a mechanicaloil pump.

Referring back to FIGS. 23A and 23B, the oil pump 2208 may include aninlet 2228 to collect lubricating fluid from the sump and an outlet 2230to deliver lubricating fluid to the various parts of the engine. In someembodiments, the rotation of the drive gear 2212 and the idle gear 2210cause the lubricating fluid from the sump to be drawn into the oil pumpthrough the inlet 2228 and forced out of the pump through the outlet2230. The inlet 2228 preferably includes a filter to remove particulatesthat may be found in the lubricating fluid prior to its being drawn intothe oil pump. In some embodiments, the inlet 2228 may be connected tothe sump via a tube, pipe, or hose. In some embodiments, the inlet 2228may be in direct fluid communication with the sump.

In some embodiments, the oil pump outlet 2230 is connected to a seriesof passageways in the various engine parts, through which thelubricating fluid is delivered to the various engine parts. The outlet2230 may be integrated with the passageways so as to be in directcommunication with the passageways, or may be connected to thepassageways via a hose or tube, or a plurality of hoses or tubes. Theseries of passageways are preferably an interconnected network ofpassageways, so that the outlet 2230 may be connected to a singlepassageway inlet and still be able to deliver lubricating fluid to theengine's lubricated parts.

FIGS. 27A-27D show one embodiments, wherein the oil pump outlet (shownas 2230 in FIG. 23B) is connected to a passageway 2700 in the rockershaft 2702 of the rocking beam drive 2704. The rocker shaft passageway2700 delivers lubricating fluid to the rocker pivot bearings 2706, andis connected to and delivers lubricating fluid to the rocking beampassageways (not shown). The rocking beam passageways deliverlubricating fluid to the connecting wrist pin bearings 2708, the linkrod bearings 2710, and the link rod passageways 2712. The link rodpassageways 2712 deliver lubricating fluid to the piston rod couplingbearing 2714. The connecting rod passageway (not shown) of theconnecting rod 2720 delivers lubricating fluid to a first crank pin 2722and the crankshaft passageway 2724 of the crankshaft 2726. Thecrankshaft passageway 2724 delivers lubricating fluid to the crankshaftjournal bearings 2728, the second crank pin bearing 2730, and the splineshaft passageway 2732. The spline shaft passageway 2732 deliverslubricating fluid to the spline shaft spline joints 2734 and 2736. Theoil pump outlet (not shown, shown in FIG. 23B as 2230) in someembodiments is connected to the main feed 2740. In some embodiments, anoil pump outlet may also be connected to and provide lubricating fluidto the coupling joint linear bearings 2738. In some embodiments, an oilpump outlet may be connected to the linear bearings 2738 via a tube orhose, or plurality of tubes or hoses. Alternatively, the link rodpassageways 2712 may deliver lubricating fluid to the linear bearings2738.

Thus, the main feed 2740 delivers lubricating fluid to the journalbearings surfaces 2728. From the journal bearing surfaces 2728, thelubricating fluid is delivered to the crankshaft main passage. Thecrankshaft main passage delivers lubricating fluid to both the splineshaft passageway 2732 and the connecting rod bearing on the crank pin2724.

Lubricating fluid is delivered back to the sump, preferably by flowingout of the aforementioned bearings and into the sump. In the sump, thelubricating fluid will be collected by the oil pump and redistributedthroughout the engine.

Tube Heat Exchanger

External combustion engines, such as, for example, Stirling cycleengines, may use tube heater heads to achieve high power. FIG. 28 is across-sectional view of a cylinder and tube heater head of anillustrative Stirling cycle engine. A typical configuration of a tubeheater head 2800, as shown in FIG. 28, uses a cage of U-shaped heatertubes 2802 surrounding a combustion chamber 2804. A cylinder 2806contains a working fluid, such as, for example, helium. The workingfluid is displaced by the piston 2808 and driven through the heatertubes 2802. A burner 2810 combusts a combination of fuel and air toproduce hot combustion gases that are used to heat the working fluidthrough the heater tubes 2802 by conduction. The heater tubes 2802connect a regenerator 2812 with the cylinder 2806. The regenerator 2812may be a matrix of material having a large ratio of surface to areavolume which serves to absorb heat from the working fluid or to heat theworking fluid during the cycles of the engine. Heater tubes 2802 providea high surface area and a high heat transfer coefficient for the flow ofthe combustion gases past the heater tubes 2802. Various embodiments oftube heater heads are discussed below, and in U.S. Pat. Nos. 6,543,215and 7,308,787, which are, as previously mentioned, incorporated byreference in their entireties.

FIG. 29 is a side view in cross section of a tube heater head and acylinder. The heater head 2906 is substantially a cylinder having oneclosed end 2920 (otherwise referred to as the cylinder head) and an openend 2922. Closed end 2920 includes a plurality of U-shaped heater tubes2904 that are disposed in a burner 3036 (shown in FIG. 30). EachU-shaped tube 2904 has an outer portion 2916 (otherwise referred toherein as an “outer heater tube”) and an inner portion 2918 (otherwisereferred to herein as an “inner heater tube”). The heater tubes 2904connect the cylinder 2902 to regenerator 2910. Cylinder 2902 is disposedinside heater head 2906 and is also typically supported by the heaterhead 2906. A piston 2924 travels along the interior of cylinder 2902. Asthe piston 2924 travels toward the closed end 2920 of the heater head2906, working fluid within the cylinder 2902 is displaced and caused toflow through the heater tubes 2924 and regenerator 2910 as illustratedby arrows 2930 and 2932 in FIG. 29. A burner flange 2908 provides anattachment surface for a burner 3036 (shown in FIG. 30) and a coolerflange 2912 provides an attachment surface for a cooler (not shown).

Referring to FIG. 30, as mentioned above, the closed end of heater head3006, including the heater tubes 3004, is disposed in a burner 3036 thatincludes a combustion chamber 3038. Hot combustion gases (otherwisereferred to herein as “exhaust gases”) in combustion chamber 3038 are indirect thermal contact with heater tubes 3004 of heater head 3006.Thermal energy is transferred by conduction from the exhaust gases tothe heater tubes 3004 and from the heater tubes 3004 to the workingfluid of the engine, typically helium. Other gases, such as nitrogen,for example, or mixtures of gases, may be used, with a preferableworking fluid having high thermal conductivity and low viscosity.Non-combustible gases are used in various embodiments. Heat istransferred from the exhaust gases to the heater tubes 3004 as theexhaust gases flow around the surfaces of the heater tubes 3004. Arrows3042 show the general radial direction of flow of the exhaust gases.Arrows 3040 show the direction of flow of the exhaust gas as it exitsfrom the burner 3036. The exhaust gases exiting from the burner 3036tend to overheat the upper part of the heater tubes 3004 (near theU-bend) because the flow of the exhaust gases is greater near the upperpart of the heater tubes than at the bottom of the heater tubes (i.e.,near the bottom of the burner 3036).

The overall efficiency of an external combustion engine is dependent inpart on the efficiency of heat transfer between the combustion gases andthe working fluid of the engine.

Returning to FIG. 29, in general, the inner heater tubes 2918 are warmerthan the outer heater tubes 2916 by several hundred degrees Celsius. Theburner power and thus the amount of heating provided to the workingfluid is therefore limited by the inner heater tube 2918 temperatures.The maximum amount of heat will be transferred to the working gas if theinner and outer heater tubes are nearly the same temperature. Generally,embodiments, as described herein, either increase the heat transfer tothe outer heater tubes or decrease the rate of heat transfer to theinner heater tubes.

FIG. 31 is a perspective view of an exhaust flow concentrator and a tubeheater head in accordance with one embodiment. Heat transfer to acylinder, such as a heater-tube, in cross-flow, is generally limited toonly the upstream half of the tube. Heat transfer on the back side (ordownstream half) of the tube, however, is nearly zero due to flowseparation and recirculation. An exhaust flow concentrator 3102 may beused to improve heat transfer from the exhaust gases to the downstreamside of the outer heater tubes by directing the flow of hot exhaustgases around the downstream side (i.e. the back side) of the outerheater tubes. As shown in FIG. 31, exhaust flow concentrator 3102 is acylinder placed outside the bank of heater tubes 3104. The exhaust flowconcentrator 3102 may be fabricated from heat resistant alloys,preferably high nickel alloys such as Inconel 600, Inconel 625,Stainless Steels 310 and 316 and more preferably Hastelloy X. Openings3106 in the exhaust flow concentrator 3102 are lined up with the outerheater tubes. The openings 3106 may be any number of shapes such as aslot, round hole, oval hole, square hole etc. In FIG. 31, the openings3106 are shown as slots. In some embodiments, the slots 3106 have awidth approximately equal to the diameter of a heater tube 3104. Theexhaust flow concentrator 3102 is preferably a distance from the outerheater tubes equivalent to one to two heater tube diameters.

FIG. 32 illustrates the flow of exhaust gases using the exhaust flowconcentrator as shown in FIG. 31. As mentioned above, heat transfer isgenerally limited to the upstream side 3210 of a heater tube 3204. Usingthe exhaust flow concentrator 3202, the exhaust gas flow is forcedthrough openings 3206 as shown by arrows 3212. Accordingly, as shown inFIG. 32, the exhaust flow concentrator 3202 increases the exhaust gasflow 3212 past the downstream side 3214 of the heater tubes 3204. Theincreased exhaust gas flow past the downstream side 3214 of the heatertubes 3204 improves the heat transfer from the exhaust gases to thedownstream side 3214 of the heater tubes 3204. This in turn increasesthe efficiency of heat transfer to the working fluid which can increasethe overall efficiency and power of the engine.

Returning to FIG. 31, the exhaust flow concentrator 3102 may alsoimprove the heat transfer to the downstream side of the heater tubes3104 by radiation. Referring to FIG. 33, given enough heat transferbetween the exhaust gases and the exhaust flow concentrator, thetemperature of the exhaust flow concentrator 3302 will approach thetemperature of the exhaust gases. In a some embodiments, the exhaustflow concentrator 3302 does not carry any load and may therefore,operate at 1000.degree. C. or higher. In contrast, the heater tubes 3304generally operate at 700.degree. C. Due to the temperature difference,the exhaust flow concentrator 3302 may then radiate thermally to themuch cooler heater tubes 3304 thereby increasing the heat transfer tothe heater tubes 3304 and the working fluid of the engine. Heat transfersurfaces (or fins) 3310 may be added to the exhaust flow concentrator3302 to increase the amount of thermal energy captured by the exhaustflow concentrator 3302 that may then be transferred to the heater tubesby radiation. Fins 3310 are coupled to the exhaust flow concentrator3302 at positions outboard of and between the openings 3306 so that theexhaust gas flow is directed along the exhaust flow concentrator,thereby reducing the radiant thermal energy lost through each opening inthe exhaust flow concentrator. The fins 3310 are preferably attached tothe exhaust flow concentrator 3302 through spot welding. Alternatively,the fins 3310 may be welded or brazed to the exhaust flow concentrator3302. The fins 3310 should be fabricated from the same material as theexhaust flow concentrator 3302 to minimize differential thermalexpansion and subsequent cracking. The fins 3310 may be fabricated fromheat resistant alloys, preferably high nickel alloys such as Inconel600, Inconel 625, Stainless Steels 310 and 316 and more preferablyHastelloy X.

As mentioned above with respect to FIG. 30, the radial flow of theexhaust gases from the burner is greatest closest to the exit of theburner (i.e., the upper U-bend of the heater tubes). This is due in partto the swirl induced in the flow of the exhaust gases and the suddenexpansion as the exhaust gases exit the burner. The high exhaust gasflow rates at the top of the heater tubes creates hot spots at the topof the heater tubes and reduces the exhaust gas flow and heat transferto the lower sections of the heater tubes. Local overheating (hot spots)may result in failure of the heater tubes and thereby the failure of theengine. FIG. 34 is a perspective view of an exhaust flow axial equalizerin accordance with an embodiment. The exhaust flow axial equalizer 3420is used to improve the distribution of the exhaust gases along thelongitudinal axis of the heater tubes 3404 as the exhaust gases flowradially out of the tube heater head. (The typical radial flow of theexhaust gases is shown in FIG. 30.) As shown in FIG. 34, the exhaustflow axial equalizer 3420 is a cylinder with openings 3422. As mentionedabove, the openings 3422 may be any number of shapes such as a slot,round hole, oval hole, square hole etc. The exhaust flow axial equalizer3420 may be fabricated from heat resistant alloys, preferably highnickel alloys including Inconel 600, Inconel 625, Stainless Steels 310and 316 and more preferably Hastelloy X.

In some embodiments, the exhaust flow axial equalizer 3420 is placedoutside of the heater tubes 3404 and an exhaust flow concentrator 3402.Alternatively, the exhaust flow axial equalizer 3420 may be used byitself (i.e., without an exhaust flow concentrator 3402) and placedoutside of the heater tubes 3404 to improve the heat transfer from theexhaust gases to the heater tubes 3404. The openings 3422 of the exhaustflow axial equalizer 3420, as shown in FIG. 34, are shaped so that theyprovide a larger opening at the bottom of the heater tubes 3404. Inother words, as shown in FIG. 34, the width of the openings 3422increases from top to bottom along the longitudinal axis of the heatertubes 3404. The increased exhaust gas flow area through the openings3422 of the exhaust flow axial equalizer 3420 near the lower portions ofthe heater tubes 3404 counteracts the tendency of the exhaust gas flowto concentrate near the top of the heater tubes 3404 and therebyequalizes the axial distribution of the radial exhaust gas flow alongthe longitudinal axis of the heater tubes 3404.

In another embodiment, as shown in FIG. 35, spacing elements 3504 may beadded to an exhaust flow concentrator 3502 to reduce the spacing betweenthe heater tubes 3506. Alternatively, the spacing elements 3504 could beadded to an exhaust flow axial equalizer 3520 (shown in FIG. 34) when itis used without the exhaust flow concentrator 3504. As shown in FIG. 35,the spacing elements 3504 are placed inboard of and between theopenings. The spacers 3504 create a narrow exhaust flow channel thatforces the exhaust gas to increase its speed past the sides of heatertubes 3506. The increased speed of the combustion gas thereby increasesthe heat transfer from the combustion gases to the heater tubes 3506. Inaddition, the spacing elements may also improve the heat transfer to theheater tubes 3506 by radiation.

FIG. 36 is a cross-sectional side view, of a tube heater head 3606 andburner 3608 in accordance with an alternative embodiment. In thisembodiment, a combustion chamber of a burner 3608 is placed inside a setof heater tubes 3604 as opposed to above the set of heater tubes 3604 asshown in FIG. 30. A perforated combustion chamber liner 3615 is placedbetween the combustion chamber and the heater tubes 3604. Perforatedcombustion chamber liner 3615 protects the inner heater tubes fromdirect impingement by the flames in the combustion chamber. Like theexhaust flow axial equalizer 3420, as described above with respect toFIG. 34, the perforated combustion chamber liner 3615 equalizes theradial exhaust gas flow along the longitudinal axis of the heater tubes3604 so that the radial exhaust gas flow across the top of the heatertubes 3604 (near the U-bend) is roughly equivalent to the radial exhaustgas flow across the bottom of the heater tubes 3604. The openings in theperforated combustion chamber liner 3615 are arranged so that thecombustion gases exiting the perforated combustion chamber liner 3615pass between the inner heater tubes 3604. Diverting the combustion gasesaway from the upstream side of the inner heater tubes 3604 will reducethe inner heater tube temperature, which in turn allows for a higherburner power and a higher engine power. An exhaust flow concentrator3602 may be placed outside of the heater tubes 3604. The exhaust flowconcentrator 3602 is described above with respect to FIGS. 31 and 32.

Another method for increasing the heat transfer from the combustion gasto the heater tubes of a tube heater head so as to transfer heat, inturn, to the working fluid of the engine is shown in FIG. 37. FIG. 37 isa perspective view of a tube heater head including flow diverter fins inaccordance with an embodiment. Flow diverter fins 3702 are used todirect the exhaust gas flow around the heater tubes 3704, including thedownstream side of the heater tubes 3704, in order to increase the heattransfer from the exhaust gas to the heater tubes 3704. Flow diverterfin 3702 is thermally connected to a heater tube 3704 along the entirelength of the flow diverter fin. Therefore, in addition to directing theflow of the exhaust gas, flow diverter fins 3702 increase the surfacearea for the transfer of heat by conduction to the heater tubes 3704,and thence to the working fluid.

FIG. 38 is a top view in cross-section of a tube heater head includingflow diverter fins in accordance with an embodiment. Typically, theouter heater tubes 3806 have a large inter-tube spacing. Therefore, someembodiments as shown in FIG. 38, the flow diverter fins 3802 are used onthe outer heater tubes 3806. In an alternative embodiment, the flowdiverter fins could be placed on the inner heater tubes 3808 (also shownin FIG. 39 as 3908). As shown in FIG. 38, a pair of flow diverter finsis connected to each outer heater tube 3806. One flow diverter fin isattached to the upstream side of the heater tube and one flow diverterfin is attached to the downstream side of the heater tube. In someembodiments, the flow diverter fins 3802 are “L” shaped in cross sectionas shown in FIG. 38. Each flow diverter fin 3802 is brazed to an outerheater tube so that the inner (or upstream) flow diverter fin of oneheater tube overlaps with the outer (or downstream) flow diverter fin ofan adjacent heater tube to form a serpentine flow channel. The path ofthe exhaust gas flow caused by the flow diverter fins is shown by arrows3814. The thickness of the flow diverter fins 3802 decreases the size ofthe exhaust gas flow channel thereby increasing the speed of the exhaustgas flow. This, in turn, results in improved heat transfer to the outerheater tubes 3806. As mentioned above, with respect to FIG. 37, the flowdiverter fins 3802 also increase the surface area of the outer heatertubes 3806 for the transfer of heat by conduction to the outer heatertubes 3806.

FIG. 39 is a cross-sectional top view of a section of the tube heaterhead of FIG. 37 in accordance with an embodiment. As mentioned above,with respect to FIG. 38, a pair of flow diverter fins 3902 is brazed toeach of the outer heater tubes 3906. In some embodiments, the flowdiverter fins 3902 are attached to an outer heater tube 3906 using anickel braze along the full length of the heater tube. Alternatively,the flow diverter fins could be brazed with other high temperaturematerials, welded or joined using other techniques known in the art thatprovide a mechanical and thermal bond between the flow diverter fin andthe heater tube.

An alternative embodiment of flow diverter fins is shown in FIG. 40.FIG. 40 is a top view of a section of a tube heater head includingsingle flow diverter fins in accordance with an embodiment. In thisembodiment, a single flow diverter fin 4002 is connected to each outerheater tube 4004. In some embodiments, the flow diverter fins 4002 areattached to an outer heater tube 4004 using a nickel braze along thefull length of the heater tube. Alternatively, the flow diverter finsmay be brazed with other high temperature materials, welded or joinedusing other techniques known in the art that provide a mechanical andthermal bond between the flow diverter fin and the heater tube. Flowdiverter fins 4002 are used to direct the exhaust gas flow around theheater tubes 4004, including the downstream side of the heater tubes4004. In order to increase the heat transfer from the exhaust gas to theheater tubes 4004, flow diverter fins 4002 are thermally connected tothe heater tube 4004. Therefore, in addition to directing the flow ofexhaust gas, flow diverter fins 4002 increase the surface area for thetransfer of heat by conduction to the heater tubes 4004, and thence tothe working fluid.

FIG. 41 is a top view in cross-section of a section of a tube heaterhead including the single flow diverter fins as shown in FIG. 40 inaccordance with an embodiment. As shown in FIG. 41, a flow diverter fin4110 is placed on the upstream side of a heater tube 4106. The diverterfin 4110 is shaped so as to maintain a constant distance from thedownstream side of the heater tube 4106 and therefore improve thetransfer of heat to the heater tube 4106. In an alternative embodiment,the flow diverter fins could be placed on the inner heater tubes 4108.

Engine performance, in terms of both power and efficiency, is highest atthe highest possible temperature of the working gas in the expansionvolume of the engine. The maximum working gas temperature, however, istypically limited by the properties of the heater head. For an externalcombustion engine with a tube heater head, the maximum temperature islimited by the metallurgical properties of the heater tubes. If theheater tubes become too hot, they may soften and fail resulting inengine shut down. Alternatively, at too high of a temperature the tubeswill be severely oxidized and fail. It is, therefore, important toengine performance to control the temperature of the heater tubes. Atemperature sensing device, such as a thermocouple, may be used tomeasure the temperature of the heater tubes. The temperature sensormounting scheme may thermally bond the sensor to the heater tube andisolate the sensor from the much hotter combustion gases. The mountingscheme should be sufficiently robust to withstand the hot oxidizingenvironment of the combustion-gas and impinging flame that occur nearthe heater tubes for the life of the heater head. One set of mountingsolutions include brazing or welding thermocouples directly to theheater tubes. The thermocouples would be mounted on the part of theheater tubes exposed to the hottest combustion gas. Other possiblemounting schemes permit the replacement of the temperature sensor. Inone embodiment, the temperature sensor is in a thermowell thermallybonded to the heater tube. In another embodiment, the mounting scheme isa mount, such as a sleeve, that mechanically holds the temperaturesensor against the heater tube.

FIG. 42 is a side view in cross section of a cylinder 4204 and a burner4210. A temperature sensor 4202 is used to monitor the temperature ofthe heater tubes and provide feedback to a fuel controller (not shown)of the engine in order to maintain the heater tubes at the desiredtemperature. In some embodiments, the heater tubes are fabricated usingInconel 625 and the desired temperature is 930.degree. C. The desiredtemperature will be different for other heater tube materials. Thetemperature sensor 4202 should be placed at the hottest, and thereforethe limiting, part of the heater tubes. Generally, the hottest part ofthe heater tubes will be the upstream side of an inner heater tube 4206near the top of the heater tube. FIG. 42 shows the placement of thetemperature sensor 4202 on the upstream side of an inner heater tube4206. In some embodiments, as shown in FIG. 42, the temperature sensor4202 is clamped to the heater tube with a strip of metal 4212 that iswelded to the heater tube in order to provide good thermal contactbetween the temperature sensor 4202 and the heater tube 4206. In oneembodiment, both the heater tubes 4206 and the metal strip 4212 may beInconel 625 or other heat resistant alloys such as Inconel 600,Stainless Steels 310 and 316 and Hastelloy X. The temperature sensor4202 should be in good thermal contact with the heater tube, otherwiseit may read too high a temperature and the engine will not produce asmuch power as possible. In an alternative embodiment, the temperaturesensor sheath may be welded directly to the heater tube.

In another embodiment, as shown in FIG. 43A-B, a temperature sensormount 4320 is created with a formed strip or sheath of a refractory orhigh temperature resistant metal such as Inconel that is bonded to theexterior of the heater tube 4310. The sensor mount sheath 4320 is formedor shaped into a channel that when attached to the heater tube creates avoid that accommodates a device. In a specific embodiment, the channelis V-shaped to accommodate the insertion of a thermal sensor such as athermocouple device. The shaped channel is then bonded to the exteriorof a heater tube 4310 as shown in FIG. 43A.

FIG. 43A shows a side view of the sensor mount sheath 4320 on the heatertube 4310, while FIG. 43B is a view along the axis of the sensor mountsheath 4320. The metal should be thin enough to form, yet thick enoughto survive for the rated life of the heater head. In some embodiments,the metal is approximately between 0.005″ and 0.020″ thick. The metalmay be bent such that the bend is along the length of the strip. This“V-channel” sheath 4320 is then affixed to the exterior of the heatertube by high temperature brazing. Prior to brazing, the sheath may betack welded in several places to insure that the sheath does not moveduring the brazing process, as shown in FIG. 43A. Preferably, the brazecompound used during brazing is typically a high nickel alloy; however,any compound which will withstand the brazing temperature will work.Alternatively the sheath may be bonded to the heater tube by electronbeam or laser welding.

Now referring to FIG. 43B, a cavity 4330 is formed by affixing thesheath to the heater tube. This cavity 4330 is formed such that it mayaccept a device such as a thermocouple. When formed and brazed, thecavity may advantageously be sized to fit the thermocouple. Preferably,the fit is such that the thermocouple is pressed against the exterior ofthe heater tube. Preferably, the sheath is thermally connected to theheater tube. If the sheath is not thermally connected to the heatertube, the sheath may not be “cooled” by the working gas. The lack ofcooling may cause the sheath to operate at or near the combustion gastemperatures, which are typically high enough to eventually burn throughany metal. Brazing the sensor mount to the heater tube leads to a goodthermal contact. Alternatively, the sensor mount sheath 4320 could becontinuously welded along both sides to provide sufficient thermalconnection.

In another embodiment, as shown in FIGS. 44A-B, a second strip of metalcan be formed to create a shield 4450 over the sensor mount 4420. Theshield 4420 may be used to improve the thermal connection between thetemperature sensor, in cavity 4430, and the heater tube 4410. The shieldinsulates the sensor mount sheath 4420 from the convective heating ofthe hot combustion gases and thus improves the thermal connection to theheater tube. Furthermore, there is preferably an insulating space 4440to help further insulate the temperature sensor from the hot combustiongases as shown in FIG. 44B.

In another specific embodiment, as shown in FIGS. 45A and 45B, thetemperature sensor mount 4520 can be a small diameter tube or sleeve4540 joined to the leading edge of the heater tube 4510. FIG. 45A showsa side view of the mount on the heater tube 4510, while FIG. 45B is aview along the axis of the tube 4540 or sleeve. The sensor tube 4540 ispreferably brazed to the heater tube with a substantial braze fillet4530. The large braze fillet 4530 will maximize the thermal bond betweenthe heater tube and the sensor mount. In another embodiment, the tube orsleeve 4540 may have a shield. As described supra, an outer shield covermay help insulate the temperature sensor mount 4520 from convective heattransfer and improve the thermal connection to the heater tube.

In an alternative embodiment of the tube heater head, the U-shapedheater tubes may be replaced with several helical wound heater tubes.Typically, fewer helical shaped heater tubes are required to achievesimilar heat transfer between the exhaust gases and the working fluid.Reducing the number of heater tubes reduces the material and fabricationcosts of the heater head. In general, a helical heater tube does notrequire the additional fabrication steps of forming and attaching fins.In addition, a helical heater tube provides fewer joints that couldfail, thus increasing the reliability of the heater head.

FIGS. 46A-46D are perspective views of a helical heater tube inaccordance some embodiments. The helical heater tube, 4602, as shown inFIG. 46A, may be formed from a single long piece of tubing by wrappingthe tubing around a mandrel to form a tight helical coil 4604. The tubeis then bent around at a right angle to create a straight return passageout of the helix 4606. The right angle may be formed before the finalhelical loop is formed so that the return can be clocked to the correctangle. FIGS. 46B and 46C show further views of the helical heater tube.FIG. 46D shows an alternative embodiment of the helical heater tube inwhich the straight return passage 4606 goes through the center of thehelical coil 4604. FIG. 47 shows a helical heater tube in accordancewith one embodiment. In FIG. 47, the helical heater tube 4702 is shapedas a double helix. The heater tube 4702 may be formed using a U-shapedtube wound to form a double helix.

FIG. 48 is a perspective view of a tube heater head with helical heatertubes (as shown in FIG. 46A) in accordance with one embodiment. Helicalheater tubes 4802 are mounted in a circular pattern o the top of aheater head 4803 to form a combustion chamber 4806 in the center of thehelical heater tubes 4802. The helical heater tubes 4802 provide asignificant amount of heat exchange surface around the outside of thecombustion chamber 4806.

FIG. 49 is a cross sectional view of a burner and a tube heater headwith helical heater tubes in accordance with some embodiments. Helicalheater tubes 4902 connect the hot end of a regenerator 4904 to acylinder 4905. The helical heater tubes 4902 are arranged to form acombustion chamber 4906 (also shown in FIG. 50 as 5006) for a burner4907 that is mounted coaxially and above the helical heater tubes 4902.Fuel and air are mixed in a throat 4908 of the burner 4907 and combustedin the combustion chamber 4906. The hot combustion (or exhaust) gasesflow, as shown by arrows 4914, across the helical heater tubes 4902,providing heat to the working fluid as it passes through the helicalheater tubes 4902.

In one embodiment, the heater head 4903 (also shown in FIG. 50 as 5003)further includes a heater tube cap 4910 at the top of each helicalcoiled heater tubes 4902 to prevent the exhaust gas from entering thehelical coil portion 4901 (also shown in FIG. 50 as 5001) of each heatertube and exiting out the top of the coil. In another embodiment, anannular shaped piece of metal covers the top of all of the helicalcoiled heater tubes. The heater tube cap 4910 prevents the flow of theexhaust gas along the heater head axis to the top of the helical heatertubes between the helical heater tubes. In one embodiment, the heatertube cap 4910 may be Inconel 625 or other heat resistant alloys such asInconel 600, Stainless Steels 310 and 316 and Hastelloy X.

In another embodiment, the top of the heater head 4903 under the helicalheater tubes 4902 is covered with a moldable ceramic paste. The ceramicpaste insulates the heater head 4903 from impingement heating by theflames in the combustion chamber 4906 as well as from the exhaust gases.In addition, the ceramic blocks the flow of the exhaust gases along theheater head axis to the bottom of the helical heater tubes 4902 eitherbetween the helical heater tubes 4902 or inside the helical coil portion4901 of each heater tube.

FIG. 50 is a top view of a tube heater head with helical heater tubes inaccordance with one embodiment. As shown in FIG. 50, the return orstraight section 5002 of each helical heater tube 5000 is advantageouslyplaced outboard of gap 5009 between adjacent helical heater tubes 5000.It is important to balance the flow of exhaust gases through the helicalheater tubes 5000 with the flow of exhaust gases through the gaps 5009between the helical heater tubes 5000. By placing the straight portion5002 of the helical heater tube outboard of the gap 5009, the pressuredrop for exhaust gas passing through the helical heater tubes isincreased, thereby forcing more of the exhaust gas through the helicalcoils where the heat transfer and heat exchange area are high. Exhaustgas that does not pass between the helical heater tubes will impinge onthe straight section 5002 of the helical heater tube, providing highheat transfer between the exhaust gases and the straight section. BothFIGS. 49 and 50 show the helical heater tubes placed as close togetheras possible to minimize the flow of exhaust gas between the helicalheater tubes and thus maximize heat transfer. In one embodiment, thehelical coiled heater tubes 4901 may be arranged so that the coils nesttogether.

Pin or Fin Heat Exchanger

Now referring to FIGS. 51A and 51B, fins or pins may alternatively beused to increase the interfacial area between the hot fluid combustionproducts and the solid heater head so as to transfer heat, in turn, tothe working fluid of the engine. Heater head 5100 may have heat transferpins 5124, here shown on the interior surface of heater head 5100, inthe space between the heater head and expansion cylinder liner 5115.Additionally, as shown in FIG. 51B in a cross section of Stirling cycleengine 5196 taken along a different diameter of expansion volume 5198from that of FIG. 51A, heat transfer pins 5130 may also be disposed onthe exterior surface of heater head 5100 so as to provide a largesurface area for the transfer of heat by conduction to heater head 5100,and thence to the working fluid, from combustion gases flowing fromcombustor 5122 past the heat transfer pins. Dashed line 5131 representsthe longitudinal axis of the expansion cylinder. FIG. 51B also showsheat transfer pins 5133 lining the interior and exterior surfaces of thetop of heater head 5100, in accordance with one embodiment.Interior-facing heat transfer pins 5124 serve to provide a large surfacearea for the transfer of heat by conduction from heater head 5100 toworking fluid displaced from expansion volume 5198 by the expansionpiston and driven through regenerator chamber 5132. Additionalembodiments of heater head 5100 are disclosed in U.S. Pat. Nos.6,381,958, and 6,966,182, which, as previously mentioned, areincorporated by reference in their entireties.

Depending on the size of heater head 5100, hundreds or thousands ofinner transfer pins 5124 and outer heat transfer pins 5130 may bedesirable.

One method for manufacturing heater head 5100 with heat transfer pins5124 and 5130 includes casting the heater head and pins (or otherprotuberances) as an integral unit. Casting methods for fabricating theheater head and pins as an integral unit include, for example,investment casting, sand casting, or die casting.

While the use of pin fins is known for improving heat transfer between asurface and a fluid, the integral casting of radial pin fins on thecylindrical heater head of a Stirling engine has not been practiced norsuggested in the art, despite the fact that casting the heater head andit's heat exchange surfaces in a single step is one of the most costeffective methods to produce a heater head. The difficulty encounteredin integral casting of radial pin fins is discussed further below. A pinfin that could be cast as part of cylindrical wall would allow theinexpensive fabrication of a highly effective heater head and/or coolerfor a Stirling engine.

Castings are made by creating negative forms of the desired part. Allforms of production casting (sand, investment and injection) involvesforming extended surfaces and details by injecting material into a moldand then removing the mold from the material leaving the desirednegative or positive form behind. Removing the mold from the materialrequires that all the extended surfaces are at least parallel. In fact,good design practice requires slight draft on these extended surfaces sothat they release cleanly. Forming radial pins on the outside or insideof a cylinder would require the molds to contain tens or hundreds ofparts that pull apart in different directions. Such a mold would be costprohibitive.

In accordance various embodiments, pins or fins may be cast onto theinside and outside surface of Stirling heat exchangers using productionsand, investment or metal injection casting methods. Referring to FIGS.52A-52D and 53D, and, first, to FIG. 52A, pins 5202 are arranged intoseveral groups 5208 of parallel pins 5202 around cylindrical wall 5210of heater head 5100, shown in cross section parallel to the central axisin FIG. 52B and in cross section transverse to the central axis, in FIG.52C. It should be noted that the technology herein described mayadvantageously be applied more generally in any other heat exchangerapplication. All the pins 5202 in each group 5208 are parallel to eachother. Only the pins 5202 in the center of the group are truly radial.The pins on the outside of the group, such as those designated bynumeral 5204 in FIGS. 52C and 53D, are angled inward from a local radiussuch as to be substantially parallel to a radial line 5212 toward thecenter of the group. In addition, the pins on the outside of the groupare preferably longer, typically by a small amount, than pins closer tothe center of the group. However, the heat transfer only changes onlyslightly from the center of the group to the outside in the embodimentdepicted in FIGS. 52A-52C, and 53D in which 5 groups 5208 of parallelpins provide approximately radial pin fins around cylinder 5210.

In the casting process in accordance with some embodiments, positive ornegative molds of each group of parallel fins are formed in a singlepiece. Several mold pieces are then assembled to form the negative formfor a sand casting. In investment mold casting, the wax positive can beformed in an injection mold with only a handful of separate parts thatpull apart in different directions. The resulting mold is formed at anacceptable cost, thereby making production of a pin fin heater headeconomically practical.

Casting of a heater head having protuberances, such as pins, extendingto the interior and exterior of a part with cylindrical walls may beachieved, in accordance with various embodiments, by investment, orlost-wax, casting, as well as by sand casting, die casting, or othercasting processes. The interior or exterior protuberances, or both, maybe integrally cast as part of the head.

While typically more cheaply accomplished than machining or assembly ofthe pin arrays, casting pin arrays may still have attendant difficultiesand substantial costs. Additionally, the casting process may result in aheater head that is less than fully densely populated with pins, thusincreasing the fraction of gases failing to collide with the heater headsurface and reducing the efficiency of heat transfer.

One embodiment of the method for populating the surfaces of heater head5100 with heat transfer pins entails fabrication of heater 5100 andarrays of heat transfer pins in separate fabrication processes. An array5250 (also shown in FIG. 53B as 5350) of heat transfer pins 5252 may becast or injection molded with panel 5254 resulting in an integralbacking panel structure shown in FIG. 52D. Pin arrays 5250, aftercasting or molding, are mounted to the inner and outer surfaces of theheater head by a high temperature braze. Thus, a more densely populatedhead with a resultant low rate of gas leakage past the pins mayadvantageously be achieved. In other embodiments, panels 5254 may besecured by various mechanical means to the heater head.

Transient liquid-phase (TLP) bonding, as described, for example, in theAerospace Structural Metals Handbook, Code 4218, p. 6 (1999) isparticularly advantageous for brazing the panels to the head, sincenickel based superalloys, typically employed for fabrication of thehead, is difficult to weld by conventional processes, and operates in ahigh stress and high temperature environment. Advantages of TLP bondingin this application are that the parts braced by TLP are effectivelywelded using the parent material and have nearly the same tensilestrength properties as integrally cast parts. TLP bonds do not remelt atelevated temperatures, whereas typical brazes will remelt at the brazingtemperature. This is of particular significance in the case ofcontinuous operation at elevated temperatures where temperatureexcursions may occur, as in the present application.

The panels 5254 of pins may be attached to the interior or exterior ofeither the heater head or the cooler by other means. In one alternativeembodiment, the panel may be mechanically attached into slots at itslateral edges. The slots are provided in dividers 5306 (described in thefollowing discussion). In another embodiment, the panels are attached tothe heater head or cooler by brazing. In yet another embodiment, thepanels are attached to the heater head or cooler by sintering the panelsto the cylindrical walls of the heater head or cooler.

Dividers 5306, as shown in FIGS. 52C, 53A, and 53B, may advantageouslyimprove the heat transfer rate of the pin fin panels. Additionally, theymay provide a convenient location for locating temperature sensors.Lastly, the dividers may advantageously provide a convenient structureto which to attach panels of pins to the heater head, in one embodiment,and a parting line for casting operations, in accordance with a furtherembodiment.

Dividers 5306 may serve to improve the thermal effectiveness of the pinfin arrays in the following manner. Referring, once again, to FIG. 52A,the rate of heat transfer for a fluid flowing through staggered pin finsis significantly higher than for fluid flowing through aligned pin fins.Fluid approaching a staggered pin array 5208 would travel at a 45-degreeangle to an axial path along the length of the cylinder, with the skewdirection designated by numeral 5214. In order to provide for improvedthermal transfer, dividers 5206, 5306 are provided, in accordance someembodiments, to force the fluid flow through the staggered array of pinfins along a path designated by numeral 5212. In addition to forcing theflow to travel axially, the dividers provide convenient interfaces andjoining planes for the casting molds described above.

In certain embodiments, individual arrays 5250, each with its associatedpanel segment 5254, comprise arcuate fractions of the circumferentialdistance around the heater head. This is apparent in the top view of theheater head assembly shown in perspective in FIG. 53A. Cylinder head5320 is shown, as is exterior surface 5302 of the heater head. Backersegments supporting arrays of heat transfer pins are not shown but areinserted, during assembly, in spaces 5304 surrounding exterior surface5302 of the heater head. Between successive heat transfer pin arraysegments are trapezoidal dividers 5306 which are baffled to block theflow of exhaust gases in a downward direction through any path otherthan past the heat transfer pins.

In one embodiment, flow dividers 5306 include structures formechanically retaining the panel segments 5254 during assembly, beforebrazing, or simply to mechanically retain the panels 5254 against heaterhead 5302.

In order to maximize engine power, the hottest part of the heater headis preferably at the highest temperature allowed, considering themetallurgical creep and tensile strength, stress, and appropriatefactors of safety. Maintaining the hottest part of the heater head atthe highest temperature requires measuring the temperature of thehottest part of the heater head. The dividers provide a convenientlocation and routing for temperature sensors on the heater had to anyaxial location along the pin fin arrays. Hot gas flow path 5313 (shownalso in FIG. 51A), is defined, on the outside, by gas flow channel cover5340. Since exhaust gases do not flow through dividers 5306, atemperature sensor, such as thermocouple 5138 (shown in FIGS. 51A and53C) is advantageously disposed in divider 5306 in order to monitor thetemperature of heater head 5100 with which the temperature sensor is inthermal contact. The position of pin arrays 5250 and temperature sensor5138 mounted within divider 5306 is shown more clearly in the view ofFIG. 53B in which the pin backer has been removed.

Temperature sensing device 5138 is preferably disposed within divider5306 as depicted in FIG. 53B. More particularly, temperature sensing tip5339 of temperature sensor 5138 is preferably located in the slotcorresponding to divider 5306 as nearly as possible to cylinder head5320 in that this area is typically the hottest part of the heater head.Alternatively, temperature sensor 5138 may be mounted directly tocylinder head 5320, however location of the sensor in the slot, asdescribed, is used in some embodiments. Engine performance, in terms ofboth power and efficiency, is highest at the highest possibletemperature, yet the maximum temperature is typically limited bymetallurgical properties. Therefore, sensor 5138 should be placed tomeasure the temperature of the hottest, and therefore the limiting, partof the heater head. Additionally, temperature sensor 5138 should beinsulated from combustion gases and walls of divider 5306 by ceramicinsulation 5342, as shown in FIG. 53C. The ceramic can also form anadhesive bond with the walls of the divider to retain the temperaturesensor in place. Electrical leads 5344 of temperature sensor 5138 shouldalso be electrically insulated.

Although the burner is designed to have circumferential symmetry, hotspots may develop on heater head 5320. Adding to the problem, the alloystypically employed for fabrication of the heater head, on account oftheir high melting point, have relatively poor thermal conductivity.Once hot spots form, they are apt to endure because the gas flow outsidethe head is axial rather than circumferential, since dividers 5306(shown in FIG. 53A) impede any circumferential flow. Additionally,heating may increase local gas viscosity thereby redirecting more flowto other channels. In order to even out the temperature distribution onthe heater head, a layer of highly thermally conductive metal, such ascopper, of thickness greater than 0.001 in. and preferably about 0.005in. is applied to interior surface 5348 of heater head 5320, bydeposition or plating, or other application method. Alternatively, asimilar coating may be applied to the exterior surface, in accordancewith another embodiment.

In order to keep the size of the Stirling cycle engine small, it isimportant to maximize the heat flux from the combustion gas through theheater head. Whereas prior art employed loops of pipe in which heattransfer to the working fluid is achieved, loops engender both lowreliability (since the loops are mechanically vulnerable) and highercost, due to the more complicated loop geometry and extra materials. Thelimiting constraint on the heat flux are the thermo-mechanicalproperties of the heater head material that must be able to withstandthe high temperatures of the combustion chamber while maintaining thestructural integrity of the pressurized head. The maximum designtemperature is determined by the hottest point on the heater head whichis typically at the top of the wall. Ideally, the entire heater wall hotsection would be at this maximum temperature, as may be controlled, forexample, by controlling the fuel flow.

As combustion gases travel past the heater head in gas flow channels5113, 5313 (shown in FIG. 51A), the gas temperature decreases as heat istransferred from the gas to the heater head. As a result, the maximumallowed heater head temperature at the top of the gas flow channel mustbe set by the material used for the heater head. The material ispreferably chosen from the family of high nickel alloys, commonly knownas super alloys, such as Inconel 600 (having a maximum temperatureT.sub.max=800.degree. C. before softening), Inconel 625(T.sub.max=900.degree. C.), Inconel 754 (T.sub.max=1080.degree. C.), orHastelloy GMR 235 (T.sub.max=935.degree. C.). The gas in gas channel5113, 5313 may cool by as much as 350.degree. C. on transit through thechannel, resulting in underheating of the bottom of the hot zone.

In accordance with some embodiments, the temperature profile of theheater wall is controlled by means of heat transfer geometry, as nowdescribed. One method for controlling the geometry is by means ofproviding a variable cross-section gas flow channel 5113, 5313 (shown inFIGS. 51A and 54A). The radial dimension (perpendicular to the wall ofthe heater head), and thus the cross-section of the channel, is large atthe top of the heater wall, thereby allowing much of the gas to bypassthe pin array at the top of the wall. The bypass allows hotter gas toreach the pin array at the bottom of the wall thereby allowing thebottom pin array to operate closer to its maximum temperature. Thetemperature gradient from the top of the heater to the bottom of the hotsection (before regenerator volume 5132, shown in FIG. 51A) has beenreduced from as much as 350.degree. C. to 100.degree. C. using avariable cross-section gas flow channel.

A second method for controlling the geometry is by varying thepopulation density and the geometry of the pin array as a function ofposition along the gas flow channel. The geometry of the pins may beadjusted by varying the height/diameter (H/D) ratio of the pins. If acasting process is used to form the pin array, the range of H/D rationsmay be limited by the process. If pin rings are used, the range of H/Dratios may be extended.

Referring now to FIGS. 53E, 53F, 54A and 54B, arrow 5402 designates thepath of heated exhaust gases past heater head 5100. Outer heat transferpins 5130 intercept the heated exhaust gases and transfer heat viaheater head 5100 and inner heat transfer pins 5124 to the working fluidthat is driven from expansion cylinder 5115 along path 5404. (Forclarity, heat transfer pins 5130 and 5124 are shown schematically inFIG. 54A. Additional heat transfer pins 5130 and 5124 had been depicted,not to scale, in the view of FIGS. 53E, 53F, and 54B.) Successive heattransfer pins 5406, 5408, and 5410, for example, present a progressivelylarger cross section to the flow of exhaust gas along path 5402. Thus,while the exhaust gas has transferred some fraction of its heat prior toarrival at the lower pins, heat is extracted there with a greaterconduction rate, thereby reducing the temperature gradient between thetop 5412 and bottom 5414 of the path of working fluid between expansionvolume 5198 and regenerator volume 5132. Typical temperatures of thesurface of expansion cylinder 5115 are indicated in FIG. 54A:850.degree. C. at the top of the cylinder, 750.degree. C. at the centerof the cylinder, and 600.degree. C. at the end of the cylinder closestto the regenerator volume.

Another method for achieving more even distribution of heat from theexhaust gases to the heater head is to create a tapered divider on theoutside diameter of the heater head by means of concentric tapered pinbacker 5146, as shown in FIG. 54A. The cross-sectional view of FIG. 54Ashows how tapered pin backer 5146 allows some of the hottest exhaust gasto bypass the pins near the top of the heater head. Pin backer 5146creates a narrowing annular gap on the outside of the pins thatprogressively forces more and more of the exhaust gases into the pinheat exchanger.

Another method for increasing the surface area of the interface betweena solid such as heater head 5100 and a fluid such as combustion gases asdiscussed above is now described with reference to FIGS. 55A-55D. Aneffect analogous to that of fabricating heat transfer pins by casting orotherwise may be obtained by punching holes 5160 into a thin annularring 5162 shown in top view in FIG. 55A and in side view in FIG. 55B.The thickness of ring 5162, which may be referred to as a ‘heat transferpin ring’ is comparable to the thickness of the heat transfer pinsdiscussed above, and is governed by the strength of the heat-conductivematerial at the high temperature of the combustion gases traversingholes 5160. The shape and disposition of holes 5160 within each ring isa matter of design for a particular application, indeed, holes 5160 maynot be surrounded by solid material. The material of rings 5162 ispreferably an oxidation-resistant metal such as Inconel 625 or HastelloyGMR 235, though other heat-conducting materials may be used. Rings 5162may be produced inexpensively by a metal stamping process. Rings 5162are then mounted and brazed, or otherwise bonded, to the outer surfaceheater head 5100, as shown with respect to outer pin rings 5164 in FIG.55C, and with respect to inner pin rings 5166 in FIG. 55D. Additionalrings may be interspersed between the pin rings to control the verticalspacing between the pins. Expansion cylinder liner 5115 is shown in theinterior of inner pin rings 5166.

Heat transfer rings 5162 may be advantageously applied to the interiorof the heater head as well as to both the exterior and interior of thecooler of a thermal cycle engine. In these applications, the rings neednot be oxidation resistant. Materials including copper and nickel arepreferably used on the interior of the heater head, while the rings forthe cooler are preferably made of one of various high thermalconductivity materials including aluminum, copper, zinc, etc.

The total cross sectional area of the heat transfer pins taken in aslice perpendicular to cylinder axis 5168 need not be constant, indeed,it is advantageously varied, as discussed in detail above, in referenceto FIG. 54.

Referring to FIGS. 56A through 56C, the interior or exterior heatexchange surfaces may also be formed from various folded fin structures5600, 5602, or 5604. The folded fin structures may be made of materialsimilar to that of the heater head pressure dome or of high thermalconductivity materials such as copper which may provide improved finefficiency. Fins fabricated from high melting-point materials such asthat of the heater head 5100 (shown in FIG. 51A) may be continuous fromthe top to the bottom of the heater head. Folded fins may be fabricatedfrom sheet metal and brazed to the interior surface of the heater head.Three folded fin configurations are shown by way of example: wavy fins5600, lanced fins 5602, and offset fins 5604. In each case, the gas flowdirection is indicated by an arrow designated by numeral 5606.

Fins formed from a dissimilar metal to that of heater head 5100 areattached in axial segments to avoid differential thermal expansion frombreaking the brazed joint between the fins and the head. The offset finconfiguration of FIG. 56C advantageously provides a superior heattransfer coefficient to that of plain fins.

The use of high thermal conductivity metal for the folded fins mayadvantageously allow the fins to be made longer, thereby improving heattransfer and reducing resistance to flow of the gas and improving engineefficiency.

Heater Head Support Ribs

The walls of the heater head must be sufficiently strong, at operatingtemperatures, to withstand the elevated pressure of the working gas. Itis typically desirable to operate Stirling cycle engines at as high aworking gas pressure as possible, thus, enabling the head to withstandhigher pressures is highly advantageous. In designing the heater head,it must be borne in mind that increasing the pressure at a givenoperating temperature typically requires increasing the heater head wallthickness in direct proportion. On the other had, thickening the heaterhead wall results in a longer thermal conduction path between theexterior heat source and the working gas.

Moreover, thermal conduction increases with heat exchanger surface area,thus thermal efficiency is increased by increasing the diameter of theheater head. Stress in the wall, however, is substantially proportionalto the diameter of the head, thus increasing the head diameter, at agiven temperature and interior gas pressure, requires increasing thewall thickness in direct proportion.

The strength considerations are tantamount at typical Stirling enginehead temperatures, in fact, they drive the maximum operatingtemperature, since, as discussed, efficiency increases with temperature.Both creep and ultimate tensile strengths of materials tend to fall offprecipitously when specified elevated temperatures are reached.Referring to FIG. 57A, the yield strength at 0.2% offset and ultimatetensile strength are shown for the GMR 235 nickel alloy in typicalrepresentation of the qualitative behavior of nickel alloys. Similarly,in FIG. 57B, it can be seen that the 0.01% per hour creep rate strengthof GMR 235 falls from 40 ksi to half as the temperature rises from1500.degree. F. to 1700.degree. F.

Some embodiments provide interior ribs (or hoops) 5800, such as thosedisclosed in U.S. Pat. Nos. 6,381,958, and 6,966,182, that enhancestructural support of heater head 5801, as shown in cross-section inFIG. 58. Ribs 5800 are characterized by an interior bore 5802. The creepstrength and rupture strength of heater head 5801 is thus determinedpredominantly by an effective thickness 5804 of the heater head and theinterior bore diameter 5802. Heat conduction through the heater head isnot limited by thickness 5804 since intervening segments 5806 of thehead are narrower and provide enhanced heat conduction. Ribs 5800 notonly relieve hoop stresses on outer wall 5808 of head 5801 butadditionally provide supplemental surface area interior to the heaterhead and thus advantageously enhance heat transfer to the working fluid.

Further advantages of providing ribs 5800 interior to the heater headinclude reducing the temperature gradient across the head wall 5808 fora given rate of heat transfer, as well as allowing operation at higherhot end working temperatures. Additionally, by reducing the stressrequirements on the outer wall, alternative materials to nickel basedsuperalloys may be used, advantageously providing superior conductivityat reduced cost.

A cross section of heater head 5801 with ribs 5800 is further shown inFIG. 59. Dashed line 5910 designates the central longitudinal axis ofthe expansion cylinder. In accordance with various embodiments expansioncylinder hot sleeve 5912 may have transverse flow diverters 5914 fordirecting the flow of working gas, represented by around 5916, aroundcircumferential ribs 5800 for enhancing heat transfer to the workinggas. The additional width h of ribs 5800 contributes to the hoopstrength of heater head 5101, whereas heat transfer is governedpredominantly by the narrower thickness t of outer heater head wall5808. In typical Stirling engine applications, while the heater headexterior may be run as hot as 1800.degree. F., ribs 5800 that providestructure strength typically run no hotter than 1300.degree. F.

Advantages of enhanced hoop strength concurrent with enhanced thermalconductivity, as discussed above with reference to FIG. 58 mayadditionally be obtained in accordance with several alternateembodiments. Referring to FIGS. 60A and 60B, cross sections are shown ofa heater head 6030, wherein tubular openings 6032 run parallel to heaterhead wall 6008. As shown in the cross sectional view of FIG. 60B, takenalong line AA, tubes 6032 allow working gas to pass down the wall,enhancing heat transfer from outside the head to the working gas.Additionally, the wall 6008 may be thicker, for the same rate of heattransfer, thus providing additional strength. Moreover, the thick wallsection 6010 (also shown in FIG. 61B as 6110) interior to passages 6032remains cooler than would otherwise be the case, providing furtheradditional strength. Heater head 6030 is preferably cast with tubularpassages 6032 which may be round in cross section or of other shapes.

FIG. 61A shows a further heater head 6140 wherein tubular openings 6132run parallel to heater head wall 6108 and are interrupted by openingsthat run out to thinner sections 6142 of the heater head wall. As shownin the cross sectional view of FIG. 62B, taken along line AA, tubes 6132allow working gas to pass down the wall, enhancing heat transfer fromoutside the head to the working gas to a degree substantially enhancedover that of the straight tube design shown in FIGS. 62A and 62B.Additionally, openings 6144 provide additional area for removal ofceramic cores used in the casting process to create such long, thinholes. Increased access to the holes allows faster chemical leaching ofthe core in the course of the manufacturing process.

FIG. 62B shows yet another heater head 6250, wherein ribs 6252 aredisposed in a helix within heater head wall 6208, thereby providing thewall with enhanced rigidity in both the circumferential and axialdirections. The working gas flows through the spiral 6254 on a pathbetween the expansion piston and the heater head, on its way to theregenerator. FIG. 62B shows a transverse cross section of the heaterhead of FIG. 62A taken along line AA. Various embodiments includeemploying a linear, or other, approximation to spiral 6254, to obtaincomparable advantages of stiffening and heat transfer.

Heater head 6250 of FIGS. 62A and 62B is preferably fabricated bycasting. A side view of core assembly 6260 for use in the castingprocess is shown in FIG. 62C. It is additionally advantageous to provideribs for internal support of the dome of the heater head and to provideadditional heat exchange on the dome, thereby cooling the inner surfaceof the dome. The complementary core structure of the dome is shown inFIG. 62D, and, in cross section, as viewed from the top, in FIG. 62D. Aperspective view of core assembly 6260 is shown in FIG. 62E.

It is to be understood that the various heater head embodiments andmethods for their manufacture described herein may be adapted tofunction in a multiple heater head configuration.

Regenerator

A regenerator is used in a Stirling cycle machine, as discussed aboveand as described in U.S. Pat. Nos. 6,591,609, and 6,862,883, to add andremove heat from the working fluid during different phases of theStirling cycle. The regenerator used in a Stirling cycle machine must becapable of high heat transfer rates which typically suggests a high heattransfer area and low flow resistance to the working fluid. Low flowresistance also contributes to the overall efficiency of the engine byreducing the energy required to pump the working fluid. Additionally, aregenerator must be fabricated in such a manner as to resist spalling orfragmentation because fragments may be entrained in the working fluidand transported to the compression or expansion cylinders and result indamage to the piston seals.

One regenerator design uses several hundred stacked metal screens. Whileexhibiting a high heat transfer surface, low flow resistance and lowspalling, metal screens may suffer the disadvantage that their cuttingand handling may generate small metal fragments that must be removedbefore assembling the regenerator. Additionally, stainless steel wovenwire mesh contributes appreciably to the cost of the Stirling cycleengine.

A three dimensional random fiber network, such as stainless steel woolor ceramic fiber, for example, may be used as the regenerator, as nowdescribed with reference to FIG. 63A. Stainless steel wool regenerator6300 advantageously provides a large surface area to volume ratio,thereby providing favorable heat transfer rates at low fluid flowfriction in a compact form. Additionally, cumbersome manufacturing stepsof cutting, cleaning and assembling large numbers of screens areadvantageously eliminated. The low mechanical strength of steel wool andthe tendency of steel wool to spall may both be overcome as nowdescribed. In some embodiments, the individual steel wires 6302 and 6304are “cross-linked” into a unitary 3D wire matrix.

The starting material for the regenerator may be fibrilose and of randomfiber form such as either steel or nickel wool. The composition of thefiber may be a glass or a ceramic or a metal such as steel, copper, orother high temperature materials. The diameter of the fiber ispreferably in the range from 10 micrometers to 1 millimeter depending onthe size of the regenerator and the properties of the metal. Thestarting material is placed into a form corresponding to the final shapeof the regenerator which is depicted in cross-section in FIG. 63B. Innercanister cylindrical wall 6320, outer canister cylindrical wall 6322,and regenerator network 6300 are shown. The density of the regeneratoris controlled by the amount of starting material placed in the form. Theform may be porous to allow fluids to pass through the form.

In some embodiments, unsintered steel wool is employed as regeneratornetwork 6300. Regenerator network 6300 is then retained within theregenerator canister by regenerator retaining screens 6324 or otherfilter, thereby comprising a “basket” which may advantageously capturesteel wool fragments.

In one embodiment, applicable to starting material that is electricallyconducting, the starting material is placed in a porous form and placedin an electrolyte bath. The starting material may be a metal, such asstainless steel, for example. An electrical connection is made with thestarting material thereby forming an electrode. Cross-linking of theindividual fibers in the starting material is accomplished byelectrically depositing a second material 6306 onto the startingmaterial. The selection of the starting material will depend on suchfactors as the particular deposition technique chosen and the chemicalcompatibility of the first and second materials, as known to one ofordinary skill in the electrochemical art. During deposition, the secondmaterial will build up on the starting material and form bridges 6308between the individual fibers of the starting material in places wherethe individual fibers are in close proximity to each other. Thedeposition is continued until the bridges have grown to a sufficientsize to hold the two individual fibers rigidly in place.

The deposition duration depends on the particular deposition process andis easily determined by one of ordinary skill in the art. After thedeposition is completed, the regenerator is removed from the bath andthe form and is cleaned.

In another embodiment the starting material is placed in a form that maybe porous or not. The form containing the starting material is placed ina furnace and is partially sintered into a unitary piece. The selectionof the sintering temperature and sintering time is easily determined byone of ordinary skill in the sintering art.

In another embodiment the starting material is placed in a porous form.The form containing the starting material is placed in a chemical bathand a second material, such as nickel, is chemically deposited to formbridges between the individual fibers.

In another embodiment the starting material is a silica glass fiberwhich is placed into a porous form. The glass fiber and form is dippedin a solution of tetraethylorthosilicate (TEOS) and ethanol so that thefiber is completely wetted by the solution. The fiber and form areremoved from the solution and allowed to drain in a humid atmosphere.The solution will form meniscoidal shapes bridging fibers in closeproximity to each other. The humidity of the atmosphere will start thehydrolysis-condensation reaction that converts the TEOS to silicaforming a cross link between the two fibers. The fiber and form may beheat treated at a temperature less than 1000° C., most preferably lessthan 600° C., to remove the reactant products and form a silica bridgebetween the fibers.

In another embodiment a ceramic slurry is deposited onto a reticulatedfoam having the shape of the regenerator. The slurry is dried on thereticulated foam and heat treated to burn off the foam and sinter theceramic. The ceramic may be composed of an oxide ceramic such ascordierite, alumina, or zirconia. The composition of the ceramic slurryand the heat treatment profile is easily specified by one of ordinaryskill in the ceramic processing art.

In yet other embodiments, knit or woven wire is employed in fabricationof a regenerator as now described with reference to FIG. 64A. Inaccordance with these embodiments, knit or woven wire tube 6401 isflattened by rollers 6402 into tape 6404, in which form it is woundabout mandrel 6406 into annular layers 6408. Stainless steel isadvantageously used for knit wire tube 6401 because of its ability towithstand elevated temperature operation, and the diameter of the wireused is typically in the range of 1-2 mils, however other materials andgauges may be used in various embodiments. Alternatively, a plurality,typically 5-10, of the stainless steel wires may be loosely wound into amulti-filament thread prior to knitting into a wire tube. This processadvantageously strengthens the resulting tube 6401. When mandrel 6406 isremoved, annular assembly 6410 may be used as a regenerator in a thermalcycle engine.

Still another embodiment is now described with reference to FIGS. 64Bthrough 64E. Knit or woven wire tube 6401, shown in its rightcylindrical form in FIG. 64B, is shown scored and partially compressedin FIG. 64C. Alternatively, the scoring may be at an angle 6414 withrespect to the central axis 6412 of the tube, as shown in FIG. 64D. Tube6401 is then axially compressed along central axis 6412 to form thebellows form 6416 shown in FIG. 64E that is then disposed as aregenerator within the regenerator volume 408 (shown in FIG. 4) of aStirling cycle engine.

It is to be understood that the various regenerator embodiments andmethods for their manufacture described herein may be adapted tofunction in a multiple cylinder configuration.

Coolant Penetrating Cold-End Pressure Vessel

Referring now to FIGS. 65A-C, various cross-sections of an engine, suchas a Stirling cycle engine, are shown in accordance with someembodiments. Engine 6500 is hermetically sealed. A crankcase 6502 servesas the cold-end pressure vessel and contains a charge gas in an interiorvolume 6504. Crankcase 6502 can be made arbitrarily strong withoutsacrificing thermal performance by using sufficiently thick steel orother structural material. A heater head 6506 serves as the hot-endpressure vessel and is preferably fabricated from a high temperaturesuper-alloy such as Inconel 625, GMR-235, etc. Heater head 6506 is usedto transfer thermal energy by conduction from an external thermal source(not shown) to the working fluid. Thermal energy may be provided fromvarious heat sources such as solar radiation or combustion gases. Forexample, a burner, as previously discussed, may be used to produce hotcombustion gases (shown as 6507 in FIG. 65B) that are used to heat theworking fluid. An expansion area of cylinder (or warm section) 6522 isdisposed inside the heater head 6506 and defines part of a working gasvolume as discussed above with respect to FIG. 1. A piston 6528 is usedto displace the working fluid contained in the expansion area ofcylinder 6522.

In accordance with an embodiment, crankcase 6502 is welded directly toheater head 6506 at joints 6508 to create a pressure vessel that can bedesigned to hold any pressure without being limited, as are otherdesigns, by the requirements of heat transfer in the cooler. In analternative embodiment, the crankcase 6502 and heater head 6506 areeither brazed or bolted together. The heater head 6506 has a flange orstep 6510 that axially constrains the heater head and transfers theaxial pressure force from the heater head 6506 to the crankcase 6502,thereby relieving the pressure force from the welded or brazed joints6508. Joints 6508 serve to seal the crankcase 6502 (or cold-end pressurevessel) and bear the bending and planar stresses. In an alternativeembodiment, the joints 6508 are mechanical joints with an elastomerseal. In yet another embodiment, step 6510 is replaced with an internalweld in addition to the exterior weld at joints 6508.

Crankcase 6502 is assembled in two pieces, an upper crankcase 6512 and alower crankcase 6516. The heater head 6506 is first joined to the uppercrankcase 6512. Second, a cooler 6520 is installed with a coolant tubing(shown as 6514 in FIG. 65B) passing through holes in the upper crankcase6512. Third, the double acting pistons 6528 and drive components(designated generally as numeral 6540 in FIGS. 65A and 65C, not shown inFIG. 65B) are installed. In one embodiment, lower crankcase 6516 isassembled in three pieces, an upper section 6513, a middle section 6515,and a lower section 6517, as shown in FIGS. 65A and 65C. Middle section6515 is may be connected to upper and lower sections 6513 and 6517 atjoints 6519 and 6521, respectively, by any mechanical means known in theart, or by welding.

The lower crankcase 6516 is then joined to the upper crankcase 6512 atjoints 6518. Preferably, the upper crankcase 6512 and the lowercrankcase 6516 are joined by welding. Alternatively, a bolted flange maybe employed (as shown in FIGS. 65B and 65C).

In some embodiments a motor/generator (shown as 6501 in FIG. 65C), suchas a PM generator, may be installed into motor/generator housing (shownas 6503 in FIG. 65C), which is attached to the lower crankcase 6516, asshown in FIG. 65C. Motor/generator housing 6503 may be attached to lowercrankcase 6516 by any mechanical means known in the art, or may bewelded to lower crankcase 6516. Motor/generator housing 6503 mayassembled in two pieces, a front section 6505, which is attached tolower crankcase 6516, and a rear section 6509, which may be welded orbolted to front section 6505. In one embodiment a seal 6511 may bepositioned between the rear section 6509 and the front section 6505 ofthe motor/generator housing 6503. In some embodiments rear section 6509is removable attached to front section 6505, which serves, among otherfunctions, to allow for easy removal and installation of motor/generator6501 during engine 6500 assembly.

In order to allow direct coupling of the heater head 6506 to the uppercrankcase 6512, the cooling function of the thermal cycle is performedby a cooler 6520 that is disposed within the crankcase 6502, therebyadvantageously reducing the pressure containment requirements placedupon the cooler. By placing the cooler 6520 within crankcase 6502, thepressure across the cooler is limited to the pressure difference betweenthe working gas in the working gas volume, and the charge gas in theinterior volume 6504 of the crankcase. The difference in pressure iscreated by the compression and expansion of the working gas, and istypically limited to a percentage of the operating pressure. In oneembodiment, the pressure difference is limited to less than 30% of theoperating pressure.

Coolant tubing 6514 advantageously has a small diameter relative to thediameter of the cooler 6520. The small diameter of the coolant passages,such as provided by coolant tubing 6514, is key to achieving high heattransfer and supporting large pressure differences. The required wallthickness to withstand or support a given pressure is proportional tothe tube or vessel diameter. The low stress on the tube walls allowsvarious materials to be used for coolant tubing 6514 including, but notlimited to, thin-walled stainless steel tubing or thicker-walled coppertubing.

An additional advantage of locating the cooler 6520 entirely within thecrankcase 6502 (or cold-end pressure vessel) volume is that any leaks ofthe working gas through the cooler 6520 will only result in a reductionof engine performance. In contrast, if the cooler were to interface withthe external ambient environment, a leak of the working gas through thecooler would render the engine useless due to loss of the working gasunless the mean pressure of working gas is maintained by an externalsource. The reduced requirement for a leak-tight cooler allows for theuse of less expensive fabrication techniques including, but not limitedto, powder metal and die casting.

Cooler 6520 is used to transfer thermal energy by conduction from theworking gas and thereby cool the working gas. A coolant, either water oranother fluid, is carried through the crankcase 6502 and the cooler 6520by coolant tubing 6514. The feedthrough of the coolant tubing 6514through upper crankcase 6512 may be sealed by a soldered or brazed jointfor copper tubes, welding, in the case of stainless steel and steeltubing, or as otherwise known in the art.

The charge gas in the interior volume 6504 may also require cooling dueto heating resulting from heat dissipated in the motor/generatorwindings, mechanical friction in the drive, the non-reversiblecompression/expansion of the charge gas, and the blow-by of hot gasesfrom the working gas volume. Cooling the charge gas in the crankcase6502 increases the power and efficiency of the engine as well as thelongevity of bearings used in the engine.

In one embodiment, an additional length of coolant tubing (shown as 6530in FIG. 65B) is disposed inside the crankcase 6502 to absorb heat fromthe charge gas in the interior volume 6504. The additional length ofcoolant tubing 6530 may include a set of extended heat transfer surfaces(shown as 6548 in FIG. 65B), such as fins, to provide additional heattransfer. As shown in FIG. 65B, the additional length of coolant tubing6530 may be attached to the coolant tubing 6514 between the crankcase6502 and the cooler 6520. In an alternative embodiment, the length ofcoolant tubing 6530 may be a separate tube with its own feedthrough ofthe crankcase 6502 that is connected to the cooling loop by hosesoutside of the crankcase 6502.

In another embodiment the extended coolant tubing 6530 may be replacedwith extended surfaces on the exterior surface of the cooler 6520 or thedrive housing (shown as 6572 in FIGS. 65A and 65C). Alternatively, a fan(shown as 6534 in FIG. 65B) may be attached to the engine crankshaft(shown as 6542 in FIG. 65C) to circulate the charge gas in interiorvolume 6504. The fan 6534 may be used separately or in conjunction withthe additional coolant tubing 6530 or the extended surfaces on thecooler 6520 or drive housing 6572 to directly cool the charge gas in theinterior volume 6504.

Preferably, coolant tubing 6514 is a continuous tube throughout theinterior volume 6504 of the crankcase and the cooler 6520.Alternatively, two pieces of tubing could be used between the crankcaseand the feedthrough ports of the cooler. One tube carries coolant fromoutside the crankcase 6502 to the cooler 6520. A second tube returns thecoolant from the cooler 6520 to the exterior of the crankcase 6502. Inanother embodiment, multiple pieces of tubing may be used between thecrankcase 6502 and the cooler in order to add tubing with extended heattransfer surfaces inside the crankcase volume 6504 or to facilitatefabrication. The tubing joints and joints between the tubing and thecooler may be brazed, soldered, welded or mechanical joints.

Various methods may be used to join coolant tubing 6514 to cooler 6520.Any known method for joining the coolant tubing 6514 to the cooler 6520may be used in various embodiments. In one embodiment, the coolanttubing 6514 may be attached to the wall of the cooler 6520 by brazing,soldering or gluing. Cooler 6520 is in the form of a cylinder placedaround the cylinder 6522 and the annular flow path of the working gasoutside of the cylinder 6522. Accordingly, the coolant tubing 6514 maybe wrapped around the interior of the cooler cylinder wall and attachedas mentioned above.

Alternative cooler configurations are presented in FIGS. 65D-65G thatreduce the complexity of the cooler body fabrication. FIG. 65D shows oneembodiment of a side view of a Stirling cycle engine including coolanttubing. In FIG. 65D, cooler 6552 includes a cooler working space 6550.Coolant tubing 6548 is placed within the cooler working space 6550, sothat the working gas can flow over an outside surface of coolant tubing6548. The working gas is confined to flow past the coolant tubing 6548by the cooler body 6552 and a cooler liner 6526. The coolant tube passesinto and out-of the working space 6550 through ports in either thecooler 6552 or the drive housing 6572 (shown in FIGS. 65A and 65C). Thecooler casting process is simplified by having a seal around coolantlines 6548. In addition, placing the coolant line 6548 in the workingspace improves the heat transfer between the working fluid and thecoolant fluid. The coolant tubing 6548 may be smooth or may haveextended heat transfer surfaces or fins on the outside of the tubing toincrease heat transfer between the working gas and the coolant tubing6548. In another embodiment, as shown in FIG. 65E, spacing elements 6554may be added to the cooler working space 6550 to force the working gasto flow closer to the coolant tubes 6548. The spacing elements areseparate from the cooler liner 6526 and the cooler body 6552 to allowinsertion of the coolant tube and spacing elements into the workingspace.

In another embodiment, as shown in FIG. 65F, coolant tubing 6548 isovercast to form an annular heat sink 6556 where the working gas canflow on both sides of the cooler body 6552. The annular heat sink 6556may also include extended heat transfer surfaces on its inner and outersurfaces 6560. The body of the cooler 6552 constrains the working gas toflow past the extended heat exchange surfaces on heat sink 6556. Theheat sink 6556 is typically a simpler part to fabricate than the cooler6520 in FIGS. 65A and 65B. The annular heat sink 6556 provides roughlydouble the heat transfer area of cooler 6520 shown in FIGS. 65A and 65B.In another embodiment, as shown in FIG. 65G, the cooler liner 6526 canbe cast over the coolant lines 6548. The cooler body 6552 constrains theworking gas to flow past the cooler liner 6562. Cooler liner 6526 mayalso include extended heat exchange surfaces on a surface 6560 toincrease heat transfer.

Returning to FIG. 65B, one method for joining coolant tubing 6514 tocooler 6520 is to overcast the cooler around the coolant tubing. Thismethod is described, with reference to FIGS. 66A and 66B, and may beapplied to a pressurized close-cycle machine as well as in otherapplications where it is advantageous to locate a cooler inside thecrankcase.

Referring to FIG. 66A, a heat exchanger, for example, a cooler 6520(shown in FIGS. 65A and 65B) may be fabricated by forming ahigh-temperature metal tubing 6602 into a desired shape. In oneembodiment, the metal tubing 6602 is formed into a coil using copper. Alower temperature (relative to the melting temperature of the tubing)casting process is then used to overcast the tubing 6602 with a highthermal conductivity material to form a gas interface 6604 (and 6532 inFIG. 65B), seals 6606 (and 6524 in FIG. 65B) to the rest of the engineand a structure to mechanically connect the drive housing 6572 (shown inFIG. 2) to the heater head 6506 (shown in FIG. 65B. In one embodiment,the high thermal conductivity material used to overcast the tubing isaluminum. Overcasting the tubing 6602 with a high thermal conductivitymetal assures a good thermal connection between the tubing and the heattransfer surfaces in contact with the working gas. A seal is createdaround the tubing 6602 where the tubing exits the open mold at 6610.This method of fabricating a heat exchanger advantageously providescooling passages in cast metal parts inexpensively.

FIG. 66B is a perspective view of a cooling assembly cast over thecooling coil of FIG. 66A. The casting process can include any of thefollowing: die casting, investment casting, or sand casting. The tubingmaterial is chosen from materials that will not melt or collapse duringthe casting process. Tubing materials include, but are not limited to,copper, stainless steel, nickel, and super-alloys such as Inconel. Thecasting material is chosen among those that melt at a relatively lowtemperature compared to the tubing. Typical casting materials includealuminum and its various alloys, and zinc and its various alloys.

The heat exchanger may also include extended heat transfer surfaces toincrease the interfacial area 6604 (and 6532 shown in FIG. 65B) betweenthe hot working gas and the heat exchanger so as to improve heattransfer between the working gas and the coolant. Extended heat transfersurfaces may be created on the working gas side of the heat exchanger6520 by machining extended surfaces on the inside surface (or gasinterface) 6604. Referring to FIG. 65B, a cooler liner 6526 (shown inFIG. 65B) may be pressed into the heat exchanger to form a gas barrieron the inner diameter of the heat exchanger. The cooler liner 6526directs the flow of the working gas past the inner surface of thecooler.

The extended heat transfer surfaces can be created by any of the methodsknown in the art. In accordance some embodiments, longitudinal grooves6704 are broached into the surface, as shown in detail in FIGS. 67A and67E. Alternatively, lateral grooves 6708 (also shown in enlarged sectionview FIG. 67B-1E) may be machined in addition to the longitudinalgrooves 6704 (also shown in enlarged section view FIG. 67A-1B) therebycreating aligned pins 6710 as shown in FIG. 67B. In some embodiments,grooves are cut at a helical angle to increase the heat exchange area.

In an alternative embodiment, the extended heat transfer surfaces on thegas interface 6604 (as shown in 66B) of the cooler are formed from metalfoam, expanded metal or other materials with high specific surface area.For example, a cylinder of metal foam may be soldered to the insidesurface of the cooler 6604. As discussed above, a cooler liner 6526(shown in FIG. 65B) may be pressed in to form a gas barrier on the innerdiameter of the metal foam. Other methods of forming and attaching heattransfer surfaces to the body of the cooler are described in U.S. Pat.No. 6,694,731, issued Feb. 24, 2004, entitled Stirling Engine ThermalSystem Improvements, which is herein incorporated by reference in itsentirety.

Additional coolant penetrating cold-end pressure vessel embodiments aredescribed in U.S. Pat. No. 7,325,399. It is to be understood that thevarious coolant penetrating cold-end pressure vessel embodimentsreferred to herein may be adapted to function in a multiple cylinderengine configuration.

Intake Manifold

Referring now to FIGS. 68-69B, an intake manifold 6899, is shown forapplication to a Stirling cycle engine or other combustion applicationin accordance with some embodiments. Various embodiments of intakemanifold 6899 are further disclosed in U.S. Pat. No. 6,381,958. Inaccordance with some embodiments, fuel is pre-mixed with air that may beheated above the fuel's auto-ignition temperature and a flame isprevented from forming until the fuel and air are well-mixed. FIG. 68shows one embodiment including an intake manifold 6899 and a combustionchamber 6810. The intake manifold 6899 has an axisymmetrical conduit6801 with an inlet 6803 for receiving air 6800. Air 6800 is pre-heatedto a temperature, typically above 900 K, which may be above theauto-ignition temperature of the fuel. Conduit 6801 conveys air 6800flowing inward radially with respect to combustion axis 6820 to aswirler 6802 disposed within the conduit 6801.

FIG. 69A shows a cross sectional view of the conduit 6801 includingswirler 6802 in accordance with some embodiments. In the embodiment ofFIG. 69A, swirler 6802 has several spiral-shaped vanes 6902 fordirecting the flow of air 6800 radially inward and imparting arotational component on the air. The diameter of the swirler section ofthe conduit decreases from the inlet 6904 to the outlet 6906 of swirler6802 as defined by the length of the swirler section conduit 6801. Thedecrease in diameter of swirler vanes 6902 increases the flow rate ofair 6800 in substantially inverse proportion to the diameter. The flowrate is increased so that it is above the flame speed of the fuel. Atoutlet 6906 of swirler 6802, fuel 6806, which in a one embodiment ispropane, is injected into the inwardly flowing air.

In some embodiments, fuel 6806 is injected by fuel injector 6804 througha series of nozzles 6900 as shown in FIG. 69B. More particularly, FIG.69B shows a cross sectional view of conduit 6801 and includes the fueljet nozzles 6900. Each of the nozzles 6900 is positioned at the exit ofthe swirler vanes 6902 and is centralized between two adjacent vanes.Nozzles 6900 are positioned in this way for increasing the efficiency ofmixing the air and fuel. Nozzles 6900 simultaneously inject the fuel6806 across the air flow 6800. Since the air flow is faster than theflame speed, a flame will not form at that point even though thetemperature of the air and fuel mixture is above the fuel'sauto-ignition temperature. In some embodiments, where propane is used,the preheat temperature, as governed by the temperature of the heaterhead, is approximately 900 K.

Referring again to FIG. 68, the air and fuel, now mixed, referred tohereafter as “air-fuel mixture” 6809, is transitioned in directionthrough a throat 6808 which has a contoured fairing 6822 and is attachedto the outlet 6807 of the conduit 6801. Fuel 6806 is supplied via fuelregulator 6824.

Throat 6808 has an inner radius 6814 and an outer dimension 6816. Thetransition of the air-fuel mixture is from a direction which issubstantially transverse and radially inward with respect to combustionaxis 6820 to a direction which is substantially parallel to thecombustion axis. The contour of the fairing 6822 of throat 6808 has theshape of an inverted bell such that the cross sectional area of throat6808 with respect to the combustion axis remains constant from the inlet6811 of the throat to outlet 6812 of the throat. The contour is smoothwithout steps and maintains the flow speed from the outlet of theswirler to the outlet of the throat 6808 to avoid separation and theresulting recirculation along any of the surfaces. The constant crosssectional area allows the air and fuel to continue to mix withoutdecreasing the flow speed and causing a pressure drop. A smooth andconstant cross section produces an efficient swirler, where swirlerefficiency refers to the fraction of static pressure drop across theswirler that is converted to swirling flow dynamic pressure. Swirlefficiencies of better than 80% may typically be achieved in practice.Thus, the parasitic power drain of the combustion air fan may beminimized.

Outlet 6812 of the throat flares outward allowing the air-fuel mixture6809 to disperse into the chamber 6810 slowing the air-fuel mixture 6809thereby localizing and containing the flame and causing a toroidal flameto form. The rotational momentum generated by the swirler 6802 producesa flame stabilizing ring vortex as well known in the art.

Referring to FIG. 70, a cross-section is shown of combustor 7022 andexhaust gas flow path 7013, as described above in reference to earlierfigures. In accordance with another embodiment it is recognized that thecombustion exhaust gases remain above the temperature of combustion ofthe fuel well beyond the region of combustor 7022, and that, since thefuel/air mixture is typically exceedingly lean, adequate oxidant remainsfor recombustion of the exhaust gases.

FIG. 70 further illustrates the use of a temperature sensor 7002,typically a thermocouple, to monitor the temperature of heater head 7020at the top of external pin array 7030 and thereby to control the fuelflow such as to maintain the temperature at sensor 7002 below atemperature at which the heater head significantly loses strength. Thetemperature at sensor 7002 is preferably maintained approximately50.degree. C. below the melting temperature of the heater head material.

In the configuration depicted in FIG. 70, the use of avariable-cross-section gas flow bypass channel 7004 is illustrated, asdescribed above. The taper of the bypass channel is greatly exaggeratedfor clarity of depiction. Even where a bypass channel is employed, thetemperature profile as a function of distance from the top of the heaterhead is not flat, as would be preferable. Two additional temperaturesensors, 7006 and 7008, are shown at the middle and bottom,respectively, of external pin array 7030, whereby the temperature of theexhaust gas may be monitored.

In accordance some embodiments, additional fuel is added to the exhaustgases at nozzle 7010 via afterburner fuel line 7012. Nozzle 7010 may bea ring burner, circumferentially surrounding heater head 7020 and facingexternal pin array 7030 between the positions designated in FIG. 70 bytemperature sensors 7002 and 7006. The fuel flow through afterburnerfuel line 7012 may be controlled on the basis of the exhaust gastemperature measured by temperature sensor 7008. The precise position oftemperature 7008 is preferably such as to measure the maximumtemperature of the external pin array produced by the combustion of fuelexiting from afterburner nozzle 7010.

Referring to FIG. 71A, a side view is shown in cross section of a burnerand heat recovery system, designated generally by numeral 7100, for athermal cycle engine in accordance some embodiments. In the embodimentshown, heat is exchanged between hot exhaust gases, heated in combustor7022, and air drawn in at air inlet 7104 in a heat exchanger 7106 thatis external to the heater head assembly. Additionally shown is fuelinlet 7108 and igniter 7110 used to initiate ignition in the combustor.Exhaust stream 7112 traverses heat transfer pins 7030 before beingchanneled to heat exchanger 7106. A seal ring 7114 of copper, or othermetal of sufficiently high melting temperature, forms a rod type seal onheater head flange 7116 just below the bottom row of heat transfer pins7030. Copper ring 7114 fits tightly on heater head flange 7116 producinga labyrinth seal. The right-hand portion of the cross-sectional view ofFIG. 71A, showing the region of the seal, is shown, enlarged, in FIG.71B. Copper seal ring 7114 fits tightly on heater head 7101 and has aclose fit within annular groove 7118 on the bottom surface of burnercover 7120. The configuration of ring 7114 in groove 7118 produces alabyrinth seal causing the exhaust gas, in exhaust plenum 7122 to travela convoluted path around the back side of seal ring 7114 therebylimiting exhaust gas leakage. The tight fit of ring 7114 onto head 7101limits exhaust gas leakage axially out of the burner.

It is to be understood that the various intake manifold embodimentsdescribed herein may be adapted to function in a multiple burnerconfiguration.

Gaseous Fuel Burner

Definitions: As used in this section of the detailed description, thefollowing terms shall have the meanings indicated, unless the contextotherwise requires: Fuel-Air Equivalence ratio (.phi.)=Actual Fuel-AirMass Ratio/Stoichiometric Fuel-Air Mass Ratio. The stoichiometricfuel-air mass ratio is defined as the mass ratio needed to balance thefuel+air chemical equation. The stoichiometric fuel-air mass ratio iswell known for common fuels such as propane (0.0638 g fuel/g air) andcalculable for gases such as biogas.

FIG. 72 shows one embodiment of the engine 7212 embodiment having agaseous fuel burner 7201. Various embodiments of the gaseous fuel burner7201 are also disclosed in U.S. patent application Ser. No. 11/122,447,filed May 5, 2005, published Nov. 10, 2005, which is herein incorporatedby reference in its entirety. This embodiment may be used in the contextof a Stirling cycle engine, however, other embodiments of the machineare not limited to such applications. Those skilled in the art willappreciate that the present machine may have application in othersystems, such as, with other types of external combustion engines.

The use of an ejector in a gaseous fuel burner advantageously can solvesome of the challenges faced by the traditional gaseous fuel burners.First, using an ejector can eliminate the need for additional equipment,controls, and space, such as, a gaseous fuel pump, fuel controlcircuitry, and the associated components. Furthermore, using an ejectorsuch as a venturi simplifies the fuel control system by eliminating theneed for a separate fuel control scheme. Based on the corresponding riseof the vacuum with the airflow, and subsequently, an increased fuelflow, the burner power can be regulated by regulating the airflow.Accordingly, removing separate fuel control simplifies the developmentand implementation of automatic burner control in a gaseous fuel burnerwith an ejector.

Secondly, the corresponding rise of the vacuum with airflow also resultsin an approximately steady fuel-air ratio despite changes in temperatureand airflow rates. The resulting steady fuel-air ratio simplifies thefuel control and operation of the burner, by eliminating the need forcomplex exhaust sensor/feedback fuel control mechanisms.

Referring to FIG. 72, a gaseous fuel burner 7201 comprises an ejector7240, a heat exchanger 7220, a combustion chamber 7250, and a blower7200 (shown as 7300 in FIG. 73A). The term ejector as used here includeseductors, siphons, or any device that can use the kinetic energy of onefluid to cause the flow of another fluid. Ejectors are a reliable way ofproducing vacuum-based fuel flow systems with low initial cost, lack ofmoving parts, and simplicity of operation.

Referring again to FIG. 72, in a some embodiments, the ejector 7240 is aventuri. The venturi 7240 is positioned downstream of the outlet of theair preheater or heat exchanger 7220, in a venturi plenum 7241 andproximal to the combustion chamber 7250. A blower 7200 forces airthrough the venturi 7240. The flow of air through the venturi draws in aproportional amount of fuel through the fuel inlet ports 7279. The fuelinlet ports 7279 are placed at the venturi throat 7244 where the throathas the lowest pressure. The ports 7279 are sized to produce plumes offuel across the airflow that promote good mixing within the venturi7240. This fuel-air mixture exits the venturi 7240 and forms aswirl-stabilized flame in the combustion chamber 7250. The venturi 7240draws in an amount of fuel that is substantially linearly proportionalto the airflow regardless of airflow rates and temperature of the airentering the venturi 7240.

In a some embodiments as shown in FIGS. 73A and 73B, placing the venturi7340 between the air preheater 7320 and the combustion chamber 7350promotes a substantially steady air-fuel ratio over a wide range ofairflows and venturi temperatures. FIG. 73A is a schematic drawing ofthe burner including the components of the burner such as a blower 7300,a preheater 7320, a venturi 7340, and fuel supply 7372. The drawing alsoincludes a load heat exchanger or heater head 7390 (also shown in FIGS.76-78 as 7290). The load heat exchanger 7390 is the heat exchanger ofthe engine or process that absorbs the thermal power of the hot gasesleaving the combustion chamber 7350 in the burner at some elevatedtemperature. The partially cooled burned gases then enter the exhaustside of the air preheater, where they are further cooled by incomingcombustion air. FIG. 73B shows the pressure map of the same componentsarranged linearly. The air pressure supplied by the blower, the fuelsupply pressure, and the ambient pressure are all indicated. The massflow rate (m′) of fuel into the burner is controlled by the differencebetween the fuel supply pressure at 7372 and the pressure in the venturithroat 7344 (shown in FIG. 72 as 7244) and the fuel temperature at thedominant restriction:

-   -   m′.sub.FUEL.varies.(P.sub.FUEL-P.sub.THROAT).sup.0.5/T.sub.FUEL.sup.0.5

The pressure in the throat (P.sub.THROAT) is set by the pressure dropthrough the exhaust side of the preheater 7320 plus the pressure dropthrough the heater head tubes 7390 minus the suction generated by theventuri throat 7344. The pressure drops 7320, 7390 and the throatsuction pressure 7344 are all proportional to the airflow rate and theventuri temperature.

-   -   P.sub.THROAT.varies.m′.sub.AIR.sup.2*T.sub.VENTURI

Combining these equations shows that the fuel flow will varyapproximately linearly with the airflow:

-   -   m′.sub.FUEL.varies.[P.sub.FUEL-(m′.sub.AIR.sup.2*T.sub.VENTURI)].sup.0.5/T-.sub.FUEL.sup.0.5

Regulating the fuel pressure to near ambient pressure, the fuel flow isapproximately linear with airflow.

-   -   m′.sub.FUEL.varies.m*.sub.AIR*(T.sub.VENTURI/T.sub.FUEL).sup.0.5        Thus, locating the dominant fuel restriction 7378 (shown as 7278        in FIG. 72) within the venturi plenum (shown as 7241 in FIG. 72)        provides for an approximately steady fuel-air ratio over a wide        range of airflow rates and venturi temperatures.    -   m′.sub.FUEL/m′.sub.AIR.varies.constant

FIG. 74 shows one embodiment of the ejector such as the venturi. In thisembodiment, the size of the opening of the venturi throat 7244determines the amount of suction present at the throat 7244. In aspecific embodiment, the venturi throat is approximately 0.24 inches indiameter. Referring back to FIGS. 72 and 74, fuel delivery means arecoupled to the venturi 7240. The fuel delivery means may be manifolds,fuel lines or fuel tubes. The fuel delivery means may include othercomponents such as a fuel restriction 7278, fuel inlet ports 7279 andfuel valves (not shown). Fuel supplied by a pressure regulator 7272flows through a manifold 7273 and fuel inlet ports 7279 into therelatively lower pressure in the throat 7244. In one embodiment the fuelinlet ports 7279 provide the largest portion of the pressure drop in thefuel delivery means. Preferably, making the fuel inlet ports the largestrestriction in the fuel delivery means assures that the restrictionoccurs at the venturi temperature and maximizes fuel-air mixing byproducing the largest possible fuel plumes. Referring back to FIG. 72,the fuel and air flow into the divergent cone or diffuser 7248 of theventuri, where static pressure is recovered. In the diffuser 7248, theentrained fuel mixes with the air to form an ignitable fuel air mixturein the combustion chamber 7250. The ignitable fuel-air mixture thenenters the combustion chamber 7250, where the igniter 7260 may ignitethe mixture, and the tangential flow induced by a swirler 7230 creates aswirl-stabilized flame. Using an ejector 7240 to draw the gaseous fuelinto the combustion chamber eliminates the need for a high-pressuregaseous fuel pump to deliver the fuel.

In one embodiment, the venturi 7240 is constructed from high temperaturematerials to withstand high temperatures and maintain its structuralintegrity. For the embodiment of FIG. 74, the dimensions of the venturican be approximately 0.9 inches diameter inlet and outlets with anapproximately 0.24 inches diameter throat. The half angles of theconvergent cone and divergent cones can be 21.degree. and 7.degree.respectively and the throat can be 0.25 inches long. In this embodiment,the venturi can be constructed from Inconel 600. Alternatively, otherhigh temperature metals could be used including, but not limited toStainless Steels 310, 316L, 409 and 439, Hastalloy C76, Hastalloy X,Inconel 625 and other super alloys.

In one embodiment, as shown in FIG. 72, a swirler 7230 is locatedupstream of the venturi 7240 and advantageously creates a tangentialflow of air through the venturi. As is well known in the art, thetangential flow from the swirler can create an annular vortex in thecombustion chamber, which stabilizes the flame. Additionally, theswirler 7230 increases the suction pressure at the venturi throat 7244by increasing the local air velocity over the fuel inlet ports 7279.Adding the swirler allows the venturi throat 7244 to be made larger fora given suction pressure. Furthermore, the swirling action induced bythe swirler 7230 can suppress fluctuations in the combustion chamberpressure from propagating upstream to the venturi 7240. Such pressurefluctuations can temporarily slow or stop the flow of fuel gas into theventuri 7240. The swirler 7230 thereby facilitates a steady fuel-airratio in the combustion chamber for steady airflows. The swirler 7230may be a radial swirler.

In other embodiments, the gaseous burner can be connected to multiplefuel sources. In this configuration, the burner may be fired, lit orignited with a type of fuel and then run with a different type of fuel.The use of multiple fuel sources may require a fuel delivery means tunedfor each fuel. FIGS. 75, 75A, and 75B show embodiments for two fuelswith significantly different energy densities such propane and naturalgas. In this embodiment, the fuel delivery means for the denser propanemust be approximately three times more restrictive than the fueldelivery means for the less dense natural gas or methane. In theembodiment shown in FIG. 75, the venturi has different manifolds andfuel ports for each fuel. High-density fuels such as propane wouldrequire the more restrictive fuel inlet ports 7279, while a low-densityfuel such as natural gas would require less restrictive fuel inlet ports7279A. This configuration retains the highest resistance to fuel flow atthe venturi temperature. However, the embodiment of the venturi in FIG.75 may be more difficult to manufacture and have a higher-pressure lossdrop due to the long narrow passage.

Another embodiment for a gaseous burner with multiple fuel sources isshown in FIG. 75A. In this embodiment, a fuel selector valve 7276directs the fuel through an additional fuel restriction such as 7278A or7278B for a dense gas or a less dense gas respectively. The multi-portvalve 7276 allows any number of predefined gases to be burned by thesame burner. Predefined gases such as natural gas, liquid petroleum gas(LPG) or biogas can be burned in the same burner by simply setting aselector valve to the corresponding fuel setting. Alternatively, otherembodiments can have multiple settings for different qualities of biogasas the carbon dioxide fraction in biogas can vary from 50% to 20%. Thefuel restrictors may be placed outside the burner as shown in FIG. 75Aor alternatively they can be located in the entrances to the manifold7273. If restrictions 7278 are placed outside of the burner, thensignificant part of the fuel-delivery-means pressure drop is not at theventuri temperature and thus the fuel-air ratio may vary with theventuri temperature. The burner will run initially leaner and getprogressively richer as the hotter faster air flowing through theventuri exerts a stronger vacuum on the fuel. In addition, moving asignificant part of the pressure drop from the fuel ports 7279, the fuelwill not penetrate as far into the air stream. Nevertheless, locatingmultiple restrictors 7278 for different gases may make the fabricationof the part easier.

An alternative embodiment, that provides significant flexibility in thefuel-air ratio control and fuel gas usages, is shown in FIG. 75B. Inthis embodiment, the two fuel sources, 7272A and 7272B are regulated totheir individual pressure and flows though separate fuel delivery meansadjusted for each fuel. Each fuel delivery means includes two or morerestrictions in parallel 7206A and 7208A, and 7206B and 7208B with oneor more valves 7202A, and 7202B, respectively, to vary the pressure dropof the fuel delivery means. The valves may be manually or automaticallyactuated. Fuel selector 7276 connects fuel delivery means to theventuri, while closing the other fuel off.

The multiple restrictions 7206A and 7208A, and 7206B and 7208B and thevalves 7202A and 7202B allow the pressure drop of the fuel deliverymeans to be adjusted during burner warm-up. Thus the fuel-air ratio canbe roughly maintained as the suction pressure increases with increasingventuri temperature. The multiple restrictions can also adjust forchanging fuel gas density. A changing fuel gas density may occur whenthe gaseous fuel burner is connected to biogas digester, wherein thebiogas digester is the source of fuel. In a biogas digester embodiment,the carbon dioxide (CO.sub.2) content and therefore the energy densitycan vary weekly. In this embodiment, if the CO.sub.2 content increases,the pressure-drop through the fuel delivery means must be reduced toallow higher flows of the less energy dense fuel gas. In addition, themultiple restrictions can improve the ignition of the fuel gas byproviding a richer fuel-air mixture for lighting. The richer mixture isprovided by opening additional valves 7202A or 7202B, which also reducesthe pressure-drop of the fuel delivery means. Once the burner is lit,the valve 7202A or 7202B may be closed to produce a leaner flame. Asdescribed supra, once the burner is lit, the burner may be run on adifferent fuel. A fuel selector may be used to switch the fuel types.Alternatively, an embodiment with a multiple fuel selector facilitatesvarying the fuel-air ratio during the operation of the burner.

Now referring to FIGS. 75A and 75B, the fuel selector 7276 may enablethe burner to be lit on one fuel and run on a different type of fuel.This can be important if one fuel is too weak to ignite, but will burnin a warmed up burner. In one example, the burner may be lit on a higherdensity fuel such as propane. Once the burner is warmed up, the fuelselector 7276 is moved to draw in a low-density biogas.

FIG. 76 depicts an embodiment where an automated controller 7288 adjustsa variable restriction 7292 such as a variable flow valve in the fueldelivery means to hold the exhaust oxygen constant as measured by awide-range lambda sensor or UEGO 7286. In this embodiment, the automatedscheme allows any fuel from biogas to propane to be connected to theburner and the control system can compensate for the changing fueldensity. In this embodiment, the automated controller can restrict thefuel path for dense fuels such as propane and open up the fuel path forlow-density fuels such as methane and biogas. Ignition would beaccomplished by starting the variable restrictor 7292 in the fully openposition, which will produce the richest mixture then closing it untilthe fuel-air mixture is ignited. After ignition, the controller cancontrol the fuel flow to achieve the desired exhaust oxygen level. It isalso envisioned that such an embodiment would allow the fuel air ratioto be adjusted during warm-up to optimize efficiency and burnerstability.

In another embodiment as shown in FIG. 77, the gaseous fuel burner is ahigh efficiency burner for an external combustion engine such as aStirling cycle engine. The burner includes manual controls to controlthe burner. The manual controls include a ball valve 7270 to manuallyselect a fuel type, a trim valve 7274 to adjust the fuel-air ratio and arheostat 7702 to control the blower speed, and subsequently the airflow.The preheated air 7222 in the venturi 7240 draws in the fuel from a fuelsource 7272. The fuel then mixes with the preheated air to create afuel-air mixture. The fuel-air mixture flows into the combustion chamber7250 where it burns. In this embodiment a microprocessor/controller 7288holds the heater head temperature constant as measured by thetemperature sensor 7289 by varying the engine speed. Furthermore, theblower-speed determines the burner power output and thus the enginepower output. In an alternative embodiment, the fuel trim valve 7274 isnot included.

Referring now to FIG. 78 the gaseous fuel burner 7201 is a highefficiency burner for an external combustion engine such as a Stirlingcycle engine. In this embodiment, the burner includes an oxygen sensor7286 located in the exhaust stream 7284 and a microprocessor/controller7288 to automatically restrict the fuel flow with the variablerestrictor 7292. Additionally, the burner includes a blower controller(shown as 7702 in FIG. 77). The blower controller 7702 can be adjustedby the microprocessor/controller 7288 to match the Stirling engine poweroutput with the load. In this embodiment, the burner temperature is heldconstant by varying the engine speed and the engine power output isautomatically adjusted by setting the blower speed. Accordingly, in thisembodiment, the burner can burn most gaseous fuels, including fuelswithout constant properties such as biogas.

In another embodiment as shown in FIG. 79, fuel is delivered directlyinto the venturi at a point proximal to the venturi throat 7244. Thisembodiment includes a swirler 7230 to accommodate the fuel deliverymeans such as a fuel line or fuel tube. The swirler 7230 is preferablyan axial swirler positioned in the venturi 7240 and upstream of theventuri throat 7244. In operation, the delivered fuel is entrained withthe motive air to form the fuel-air mixture. The exemplary manual orautomatic control mechanisms are adaptable to this alternate fueldelivery embodiment.

Referring back to FIG. 74, the gaseous fuel burner further comprises anigniter 7260 and a flame-monitoring device 7210. Preferably, the igniter7260 is an excitable hot surface igniter that may reach temperaturesgreater than 1150.degree. C. Alternatively, the igniter 7260 may be aceramic hot surface igniter or an excitable glow pin.

With continuing reference to FIG. 74, other embodiments include aflame-monitoring device 7210. The flame-monitoring device 7210 providesa signal in the presence of a flame. For the safe operation of the anyburner, it is important that the fuel be shut-off in the event of aflameout. The monitoring device for flame sensing is the flamerectification method using a control circuit and a flame rod.

Flame rectification, well known in the art, is one flame sensingapproach for the small, high efficiency gas burners. The device uses asingle flame rod to detect the flame. The flame rod is relativelysmaller than the grounded heater head and it is positioned within thecombustion flame. In this flame rectification embodiment, the controlunit electronics are manufactured by Kidde-Fenwal, Inc., and the flamerod is commercially available from International Ceramics and HeatingSystems

Preferably, the flame-monitoring device uses the hot surface igniter asthe flame rod. Alternatively, the flame-monitoring device may be eitherremote from the hot surface igniter, or packaged with the igniter as asingle unit.

Alternatively, an optical sensor may be used to detect the presence of aflame. A preferred sensor is an ultraviolet sensor with a clear view ofthe flame brush through an ultraviolet transparent glass and a sighttube.

It is to be understood that the various fuel burner embodimentsdescribed herein may be adapted to function in a multiple burnerconfiguration.

Fuel Pump

In accordance with some embodiments, a fuel flow to a pressurizedcombustion chamber of an engine, such as a Stirling engine, may bemetered by varying the operating parameters of a fuel pump. Variousembodiments of the fuel pump are described below and in U.S. Pat. No.7,111,460, issued Sep. 26, 2006, to Jensen et al., and U.S. patentapplication Ser. No. 11/534,979, filed Sep. 25, 2006, published Feb. 8,2007, which are herein incorporated by reference in their entireties.Desired performance may be achieved without the throttle plates orvalves or other restrictive devices that are normally used to meter thefuel flow to the combustion chamber.

FIG. 80 shows a metering pump system providing gaseous fuel to apressurized combustion chamber 8058 of an engine 8022 according to oneembodiment. A gas train, labeled generally as 8005, includes a fuel pump8014, interconnecting lines 8038, 8042 and may include a pressureregulator 8018. The fuel pump 8014 raises the fuel pressure in line 8038to a higher pressure in line 8042. The gas train delivers fuel from thegas supply to the burner 8010, where it is mixed with air and burned ina combustion chamber 8058. The fuel pump is controlled by a controller8034 that modulates the fuel flow rate by varying one or more parametersof an electrical signal sent to the fuel pump 8014. The controller mayalso regulate a blower 8060 that provides air to the combustion chamber8058 and may receive signals from sensors that report engine-operatingparameters.

In some embodiments the delivered fuel pressure in line 8038 is 6 to 13inches water column for liquefied petroleum gas. Natural gas may besupplied in line 8038 at even lower pressures of 3 to 8 inches watercolumn. Alternatively, pressure regulator 8018 can supply the fuel atlower pressures, even negative pressures. Typical fuel pressures in line8042 may range from 0.5 to 5 PSIG.

In some embodiments, fuel pump 8014 is a linear piston pump. A linearpiston pump is shown in FIG. 81. The pump includes a cylinder 8100, apiston 8102, a winding 8104, a spring 8106 and check valves 8108, 8112.When an electrical signal is applied to winding 8104, the winding pullsthe ferrous metal piston 8102 to the left, compressing the spring 8106.Check valve 8108 in the piston allows fuel to flow into compressionvolume 8110. When the electrical signal is turned off and theelectromagnetic force on the piston begins to decrease, the piston 8102is forced to the right by the spring 8106. Gas is forced out check valve8112 into the receiver volume 8114 at a higher pressure.

The flow rate of the pump can be modulated by varying the stroke of thepiston 8102. In one embodiment the signal from the controller to thepump is a half-wave alternating current (“AC”) signal, as shown in FIG.82. Circuitry to produce this signal is well known in the art. Thepiston stroke and, thus, the flow rate increases as the amplitude of theAC signal increases. In some embodiments, low amplitude signals arebiased slightly higher to improve repeatability and linearity of flowversus the driving signal. The force applied to the piston 8102 by thewindings 8104 is inversely proportional to the distance from thewindings to the piston. At low signal levels, the piston does not getvery close to the windings and small changes in the friction and inertiaof the piston will produce significant changes in the resulting pistonstroke and flow. A bias voltage is applied to bring the resting-positionof the piston closer to the windings, so that small changes in thecontroller signal that drives the piston dominate the frictional forcesand the inertia of the piston. For example, the bias voltage added tothe signal is highest at the lowest driving signal (10% signal in FIG.82) and may drop to zero before the drive signal reaches 50%. The biasis reduced at higher flow levels to take advantage of the full pumpstroke.

In another embodiment, the controller signal that drives the pump is apulse-width-modulated (“PWM”) direct current (“DC”) voltage signal. FIG.83 shows an exemplary DC waveform that may be used to drive the pump.Circuitry to generate the PWM DC signal in FIG. 83 is well known in theart. Three different drive signals are plotted versus time. These signalmodulations correspond to 10%, 50% and 90% duty cycles, which are shownfor purposes of illustration and not for limitation. Applying therectangular wave voltages of FIG. 83 to the windings 8104 of FIG. 81will cause the piston 8102 to move to the left and compress the spring8106. The stroke and, therefore, the flow will be roughly proportionalto the voltage times the duration of the signal. The lower signals, 10%and 50%, include bias voltages between signal pulses. As in the case ofthe AC drive signal, the bias voltage moves the piston closer to thewindings to provide greater piston response to small changes in thesignal and overcome the frictional and inertia forces of the piston.This bias voltage may be varied with the duration of the drive signal.The bias voltage is highest at the minimum drive signal duration and maydrop to zero before the drive voltage pulse duty cycle reaches 50%.

Other embodiments may use different controller signal waveforms to drivethe piston. In another embodiment, the piston pump of FIG. 81 can bedriven without the bias voltages shown in FIGS. 82 and 83.

In another embodiment both the frequency and the duration of the PWM DCcontroller signal modulating the pump can be varied to linearize theflow through the pump with changes in the driving signal.

In further embodiments, pump 8014 is a diaphragm pump as shown in FIG.84. In the diaphragm pump, one or more solenoidal coils 8200 drive theshaft of the pump 8202 back and forth. The shaft 8202 deflects twodiaphragms 8204 that alternatively pull gas into the chambers 8212 andthen expel it. The two wire coil is driven with an AC signal connectedto wires (8234, 8236) that drives the piston 8202 back and forth byreversing the flow of current through the coil 8200. The solenoid has apermanent magnet so that a reversing magnetic field can drive thesolenoid in opposite directions. The pumping force on the two chambers8212 is phased 180 degrees apart so that as one chamber is filled, thecompanion chamber is emptied. Check valves 8208 upstream of the pumpingchambers 8212 allow gas flow in, while the downstream valves 8210 allowflow out of the chambers and into the receiver volume 8216. Thesolenoidal coil 8200 can be driven with a full wave AC signal. Insimilar fashion to the piston pump, varying the amplitude of the ACsignal will vary the stroke and, therefore, the fuel flow through thediaphragm pump.

In another embodiment, the electrical coil 8200 in the diaphragm pump8014 of FIG. 84 can be center-tapped by adding a third wire 8232 to thecenter of the coil 8200. Wires (8234 and 8236) connect to each end ofthe coil. This three wire connection allows the piston 8202 to be drivenback and forth with a DC source. The DC source connects to the centerwire 8232 and the other connecting wires (8234 and 8236) are alternatelyconnected to ground or a negative voltage, causing current to flow inone half-coil or the other.

A three-wire coil 8302 and devices (8304, 8306, 8308) to control the DCcurrent flow to the coil are shown schematically in FIG. 85. The coilmay be used to drive a diaphragm pump solenoid, as in FIG. 85. Devices(8304, 8306, 8308) may be relays, field effect transistors (“FET”),bipolar transistors or other similar devices. The controller can varythe flow of fuel through the diaphragm pump by varying the amplitude ofapplied DC voltage signal 8312 using device 8304. Devices 8306, 8308 canbe driven as shown in FIG. 86A, where first one device is closed, thenopened and then the other device is closed and then opened. The verticalaxis of the figure corresponds to a normalized driving voltage, where asignal equal to “1” means a device is closed (i.e., shorted). Controlstrategies using PWM signals, as illustrated in FIG. 83, albeit withoutthe bias described previously for the piston pump and with suitablephasing, can be applied to each of devices 8306, 8308 in FIG. 85.

In another embodiment the amplitude and frequency of the diaphragm pumpstroke of FIG. 84 can be controlled using the three devices (8302, 8304,8306) shown in FIG. 85. The amplitude of the pump stroke is controlledby the average voltage at wire 8312. This voltage can be modulated byfast pulse-width-modulating device 8304. The stroke frequency may becontrolled as before by devices 8306 and 8308. Alternatively, device8304 can be eliminated and switches 8306 and 8308 can be pulse-widthmodulated at a high frequency during their “on” state, as illustrated inFIG. 86B. In other embodiments the center-tapped coil can be replaced bya full bridge or a half-bridge, as known to those skilled in the art.

In other embodiments for use in applications where a constant flow offuel is important, a filter 8701 may be added between pump 8700 andburner head 8706, where the fuel is mixed with the combustion air, asshown in FIG. 87A. One embodiment of the filter 8701 is an RC filtercomprising a capacitance (volume) 8702 and an orifice 8704. The volumeand orifice are sized to allow the required fuel flow and reducefluctuations in flow to a desired level. Mathematical techniques thatare well known in the art may be used to determine these filterparameters.

An acoustic filter using a volume and an orifice restrictor has theelectrical circuit analog shown in FIG. 87B. The analog of gas flow iselectrical current, the analog of gas pressure is electrical voltage,the analog of volume is electrical capacitance, the analog of flowresistance is electrical resistance and the analog of gas inertia iselectrical inductance. The orifice restrictor does not translatedirectly into this model because the orifice flow resistance isproportional to the gas flow squared (non-linear) instead of beingproportional to the gas flow as the model suggests. The model can beused through the process of linearization of flow resistance for smallsignals. The pump gas flow ripple is attenuated by the factor of1/(1+2.pi.fRC). Where “f” is the frequency component of the gas flowentering the filter from the pump. Due to the orifice restrictornon-linear characteristics, the acoustic filter has a lower attenuationat low flow causing a high burner flow ripple as a percentage of averageflow. The higher ripple can cause flame instability and higher emissionsof pollutants. This non-linearity also causes a high resistance toaverage gas flow at the higher flow rates reducing the pump maximum flowcapability.

The addition of a long thin tube 8703 to the acoustic filter providesripple attenuation through the gas mass acceleration, as shown in FIG.87C. The diagram for the electrical analog is shown in FIG. 87D. Thepump gas flow ripple is attenuated by the factor of1/[1+(LC)(2.pi.f).sup.2]. Since L and C are not a function of flow, thefilter attenuation is not affected by the flow rate and does not havethe disadvantages of the filter of FIG. 87A. Attenuation of the ripplealso increases the pump's flow rate.

Referring again to FIG. 80, in another embodiment, controller 8034modulates the output of the fuel pump 8014 to control the temperature ofthe heater tubes 8026 of the engine. The temperature of the heater tube8026 may be measured with a temperature sensor 8054, such as athermocouple, that is attached to a heater tube 8026. When the engineincreases speed, the engine draws more thermal energy from the heatertubes 8026. The tubes cool and the thermocouple 8054 reports thistemperature drop to the controller 8034, which in turn increases thefuel flow until the measured temperature is restored to a specifiedlevel. Any of the devices and methods for metering the fuel through thefuel pump, as described above, may be employed in this embodiment of themachine. Various fuel pump types including rotary vane pumps,piezoelectric pumps, crank driven piston pumps, etc., may be employed.In other embodiments, various operating parameters of a system, of whichthe pressurized chamber is a part, may be controlled by controlling thefuel pump to meter the fuel flow to the chamber. For example, the speedof an internal combustion engine or the power output of an engine may bedetermined by the controller. Alternatively, a fuel/air mixture ratio toa burner may be maintained by the controller.

It is to be understood that the various fuel pump embodiments describedherein may be adapted to function in a multiple burner configuration.

Single Burner Multiple Piston Engine

Referring now to FIGS. 88, 89A-89C, various embodiments is shown whereinan engine 8800, such as a Stirling cycle engine, having a rocking beamdrive 8802 (also shown as 810 and 812 in FIG. 8) and a plurality ofpistons (also shown in FIG. 8 as 802, 804, 806, and 808), includes asingle burner (shown as 8900 in FIGS. 89A and 89B) to heat heater heads8804 of the pistons. Heater heads 8804 may be one of the variousembodiments disclosed in the preceding section, including, but notlimited to, tube heater heads, as designated by numeral 8902 in FIG. 89A(also shown as 9116 in FIGS. 91C and 91D), or pin or fin heater heads,as designated by numeral 8904 in FIG. 89C (and also shown as 5100 inFIGS. 53D through 53F). FIG. 89B included a pin heater head 8904 havinga heater head lining 8926 fitted around the heater head 8904. Burner8900 may be one of any of the various embodiments disclosed in thepreceding sections and in U.S. Pat. No. 6,971,235, issued Dec. 6, 2005,to Langenfeld et al., which is herein incorporated by reference in itsentirety.

In one embodiment a combustion chamber 8906 is positioned above theheater heads 8900, as shown in FIGS. 89A-89C. A prechamber 8901 mayconnect the combustion chamber 8906 to a burner head 8903 via aprechamber nozzle 8908, wherein prechamber nozzle 502 may be a simplenozzle, a swirler nozzle, or a pressure swirl nozzle. The burner head8903 may house a UV window 8910 for flame detection, a fuel injector8912, which may be an air-assist fuel injector such as a Delevan siphonnozzle, and a hot surface igniter 8914. Also connected to the burnerhead 8903 are a first inlet 8916 and a second inlet 8918. One of theseinlets may be a liquid fuel inlet, and the other inlet may be anatomizing inlet.

The prechamber 8901 is a centrally located fuel preparation stagelocated upstream from the combustion chamber 8906. The prechamber 8901is where the fuel is ignited to form a diffusion flame. In oneembodiment where liquid fuel is used, the liquid fuel passes through thefirst inlet 8916. Atomizer passes through the second inlet 8918 toatomize the liquid fuel and mix with the liquid fuel in the prechamber8901. As the atomizer and liquid fuel enter the prechamber 8901 via fuelinjector 8912, it is ignited by the hot surface igniter 8914. Air mayalso pass through an intake 8920 and be preheated by a preheater 8922before it travels into the prechamber 8901, where it will mix with theatomizer and the liquid fuel. Once the mixture is preheated and formedinto a diffusion flame, it travels through the prechamber nozzle 8908into the combustion chamber 8906 to form a PPV (premixed prevaporized)flame. When the diffusion flame leaves the prechamber 8901, evaporationmay occur in the prechamber 8901 which may allow the diffusion flame tobe relit more easily, should it get flamed out or burned out.

Once the flame is in the combustion chamber 8901, the heat from theflame is used to heat the heater heads 8804. The heated gas from thecombustion chamber 8901 evenly flows over the surface of each of theheater heads 8804, wherein heater heads 8804 transfer the heat containedin the heated gas to a working fluid contained in the working space(shown as 8806 in FIG. 88) of the engine (shown as 8800 in FIG. 88). Thecombustion chamber 8901 may have apertures 8924 in its surface tofurther assist in distributing the PPV flame evenly across each of theheater heads 8804.

As described above in the current and preceding sections, the heaterheads 8804 may be a pin heater head, a folded fin heater head, or may beheater tubes. In an embodiment using a pin or fin heater head, theheater head may include a heater head lining 8926 as shown in FIG. 89B(and also shown as 5340 in FIG. 53A). The heater head lining 8926 may bea sleeve that is fitted around the heater head 8904 or it may be asleeve that is heated and expanded and then fit around the heater headsuch that when the sleeve cools it contracts and creates a snug fitaround the heater head. The heater head lining 8926 ensures uniform flowof the heated gas. Uniform flow prevents uneven temperature distributionaround the heater heads 8804 and ensures thermal efficiency, asdiscussed in detail in the preceding sections. Resultant exhaust fromthe burner may exit the burner through an exhaust 8928.

Because the burner may reach very high temperatures, the metal sued toform the burner may expand. Expansion of certain burner surfaces 8930may interfere with the efficiency of the engine or may damage the heaterheads 8804. In an alternative embodiment a compliant member may bepositioned between the heater heads 8804, or, should it be used, theheater head lining 8926 and the burner surface 8930. The compliantmember acts as a buffer against the expanding metal burner surface 8930so that the burner surface 8930 does not expand into the heater heads8804.

In an alternative embodiment a gaseous fuel, such as propane may beused. In such an embodiment the burner may include a burner head 8903and a combustion chamber 8906. The burner head 8903 may house the UVwindow 8910 for flame detection, a fuel injector 8912, which may be anair-assist fuel injector such as a Delevan siphon nozzle, and a hotsurface igniter 8914. The gaseous fuel may enter the combustion chamber8906 via the fuel injector 8912. Upon exiting the fuel injector 8912,the gaseous fuel would be ignited by the hot surface igniter 8914,thereby creating a flame inside the combustion chamber 8906. Combustionof gaseous fuels is described in detail in the preceding sections.

In yet another embodiment burner 8900 may use both gaseous and liquidfuels. Similar to the exemplary embodiment described earlier, andvarious other embodiments described in preceding sections, the burner8900 would include a combustion chamber 8906, a prechamber 8901, and aburner head 8903. The combustion chamber 8906 may be positioned abovethe heater heads 8804. A prechamber 8901 may connect the combustionchamber 8906 to a burner head 8903 via a prechamber nozzle 8908, whereinprechamber nozzle 8908 may be a simple nozzle, a swirler nozzle, or apressure swirl nozzle. The burner head 8903 may house a UV window 8910for flame detection, a fuel injector 8912, which may be an air-assistfuel injector such as a Delevan siphon nozzle, and a hot surface igniter8914. Also connected to the burner head 8903 are a first inlet 8916 anda second inlet 8918. One of these inlets may be a liquid fuel inlet andthe other inlet may be an atomizing inlet. A switch may be positionedbetween the first inlet 8916 and the second inlet 8918 so that whengaseous fuel is used, the gaseous fuel would flow through the secondinlet 8918, instead of the atomizer as described above. When liquid fuelis used, the switch would be configured such that liquid fuel would flowthrough the first inlet 8916 and atomizer would flow through the secondinlet 8918.

In a further embodiment of the burner, a blower may be coupled to burner8900.

Multiple Burner Multiple Piston Engine

Referring now to FIGS. 90 through 91B, another embodiment is shownwherein each heater head 9002 of engine 9000 may be heated by anindividual burner 9004, as shown in FIG. 90. Heater heads 9002 may beany of the various embodiments described in the preceding sections,including, but not limited to, tube heater heads, as designated bynumeral 9116 in FIGS. 91B-91D, or pin or fin heater heads, as designatedby numeral 9118 in FIG. 91A (and also shown as 5100 in FIGS. 53D through53F). Burner 9004 may be any one of the various embodiments disclosed inthe preceding sections and in U.S. Pat. No. 6,971,235.

Each burner 9004 includes a burner head 9100. Similar to previousdisclosed burner embodiments, the burner head 9100 has an igniter 9101,a fuel injector 9108, and a UV window (shown as 9107 in FIG. 91B) forflame detection. Fuel passes through a first inlet 9106, where it isheated by the igniter 9101 and formed into a flame. Preheated air,heated by the preheater 9102, may be mixed with the fuel in thecombustion chamber 9103. The heated fuel mixture forms a flame insidethe combustion chamber 9103 and heats the heater head 9002. Any exhaustfrom the burner may exit the burner via an exhaust 9105. In analternative embodiment of the burner, an atomizer may be combined withthe fuel via a second inlet 9110. In another embodiment of the burner, ablower may be incorporated to maintain an average air ration amongst theindividual burners 9004.

Yet another embodiment may include a prechamber 9111, as shown in FIG.91B. In this embodiment, the burner may include a combustion chamber9103, a prechamber 9111, and a burner head 9100. Combustion chambers9103 may be positioned above the heater heads 9002. A prechamber 9111may connect the combustion chamber 9103 to a burner head 9100 via aprechamber nozzle 9112, such as a simple nozzle, a swirler nozzle, or apressure swirl nozzle. The burner head 9100 may house the UV window 9107for flame detection, a fuel injection 9108, which may be an air-assistfuel injector such as a Delevan siphon nozzle, and a hot surface igniter9101. Also connected to the burner head 9100 are a first inlet 9106 anda second inlet 9110. One of these inlets may be a liquid fuel inlet andthe other inlet may be an atomizing inlet.

The prechamber 9111 is a centrally located fuel preparation stagelocated upstream from the combustion chamber 9103. The prechamber 9111is where the fuel is ignited to form a diffusion flame. In oneembodiment, where liquid fuel is used, the liquid fuel passes throughthe first inlet 9106. Atomizer passes through the second inlet 9110 toatomize the liquid fuel and mix with the liquid fuel in the prechamber9111. As the atomizer and liquid fuel enter the prechamber 9111 via fuelinjector 9108, it is ignited by the hot surface igniter 9101. Air mayalso pass through an intake and be preheated by a preheater 9102 beforeit travels into the prechamber 9111, where it will mix with the atomizerand the liquid fuel. Once the mixture is preheated and formed into adiffusion flame, it travels through the prechamber nozzle 9112 into thecombustion chamber 9103 to form a PPV (premixed prevaporized) flame.When the diffusion flame leaves the prechamber 9111, evaporation mayoccur in the prechamber 9111 which may allow the diffusion flame to berelit more easily, should it get flamed out or burned out.

Once the flame is in the combustion chamber 9103, the heat from theflame is used to heat the heater heads 9002. The heated gas from thecombustion chamber 9103 evenly flows over the surface of each of theheater heads 9002, wherein heater heads 9002 transfer the heat containedin the heated gas to a working fluid contained in the working space ofthe engine (shown as 9000 in FIG. 90). The combustion chamber 9103 mayhave apertures (shown as 9114 in FIG. 91A) in its surface to furtherassist in distributing the PPV flame evenly across each of the heaterheads 8804.

The principles of the present invention may be applied to all types ofengines, include Stirling engines, and may be applied to other pistonmachines utilizing cylinders such as internal combustion engines,compressors, and refrigerators. However, the present invention may notbe limited to the double-acting four-cylinder Stirling engine.

Referring now to FIG. 92, a cross section of an engine 9200 is shown.The engine 9200 is similar to the one described above with respect toFIG. 4, however, includes another embodiment of the rocking drivemechanism. The engine 9200 shown in FIG. 92 includes a rocking drivemechanism including link rods 9210, 9210, a rocking beam 9214, a rockingpivot 9224, a connecting rod 9216, a connecting pivot 9218, end pivots9220, 9222, and a crankpin 9226. Although this engine 9200 is an exampleof another embodiment of the rocking drive mechanism as discussed above,the components function in a similar fashion however, this embodimentincludes a number of additional benefits.

The configuration of the connecting rod 9216, rocking beam 9214, andconnecting pivot 9218 limit the loads on the connecting rod. Thisconfiguration additionally allows for the use of larger bearings,including standard sized tri-metal bearings. Additionally, the increaseddistance between the rocking pivot 9224 and the connecting pivot 9218increases the mechanical advantage of the rocking pivot 9224, thusreducing the loads on the connecting rod bearings

In this embodiment of the engine 9200, the side loads on the link rods9210, 9210 has been increased. However, as discussed above, the engine9200 is an oil lubricated engine, thus, concern with limiting the sideloads on the link rods 9210, 9210 has been reduced. Thus, in theembodiment shown in, for example, FIG. 4, the link rods are longer andthe loads on the connecting rod are higher. In the embodiment shown inFIG. 92, the link rods 9210, 9210 are shorter and the load on theconnecting rod 9216 is decreased.

In some embodiments, the oil pump is a Gerotor pump driven by thecrankshaft through a spline connection. In some embodiments, the oilpump is driven by the crankshaft by a gear.

With respect to the heater tubes discussed above, in some embodiments,the heater tubes include an insert. The inserts increase the efficiencyof the heat transfer between the outer wall and the inert gas. As hotgas flows around the heater tube, the insert pushes the inert gas, e.g.,helium, to the wall of the heater tube. This increases the heat transferby reducing the distance and increasing the speed of the helium. Thisalso allows for better mixing of the fluids.

In some embodiments, the inset may act as an extended heat transfersurface similar to a fin. In some embodiments, the inserts may be indirect contact with the inner wall of the heater tubes. However, in someembodiments, the inserts may not directly touch the inner wall of theheater tube but may be in close relation to the inner wall of the heatertube. Less space between the insert and the tube wall increases theefficiency of the heat exchange. In some embodiments, the rods may bepreformed with the heater tubes (i.e., through casting techniques, etc).The inserts also may provide increased heat transfer through transfer ofheat from the tube wall, through the insert and into the dead volume toheat the helium gas in the tube. The insert occupies volume in theheater tube, i.e., decreases dead volume in the heater tube.

In manufacture, after forming the heater tubes either with the insertsalready in the tubes, or, subsequent introduction of the rods into theheater tubes, in the process of bending the heater tubes, the insertsmay eliminate or mitigate kinking which may be problematic inmanufacture of the heater heads without inserts in the heater tubes.

Referring to FIGS. 93A-B, a heater tube with an insert is shown. Theinsert may take any shape, including but not limited to a rod, and inthe exemplary embodiment, is made from gear stock (see FIGS. 95A-B). Insome embodiments, as shown in FIGS. 94A-B, a helical twist may be putonto the gear stock. This helical twist exposes the gas to all sides ofthe heater tube which may increase the heat transfer or provide betteroverall heat transfer. The insert may be made of any material that willnot melt, and as the inserts, in the exemplary embodiment, will only beexposed to inert gases, corrosion is not a factor. Thus, the rods may bemade from any, but not limited to, the following materials: bronze,steel, copper, i.e., less expensive materials than the tube walls sincethe tubes will be exposed to non-inert gases. In the exemplaryembodiment, steel is used.

Referring now to FIG. 96A, an embodiment of a Stirling cycle machine isshown in cross-section and designated generally by numeral 9600. Whilethe Stirling cycle machine 9600 will be described generally withreference to the embodiment shown in FIGS. 96A and 96B, it is to beunderstood that many types of machines and engines, including but notlimited to refrigerators and compressors may similarly benefit fromvarious embodiments and improvements which are described herein,including but not limited to, external combustion engines and internalcombustion engines. In particular, the present embodiment of theStirling cycle machine is directed to improving the efficiency andoperation of a 10 Kilowatt (Kw) Stirling cycle machine, although anyother power output level are certainly contemplated and encompassedwithin the following disclosure of a machine or engine, that achieveshigh efficiency, long durability and low cost targets based onsimultaneously utilizing optimized mechanical and operational controlsystems with existing Stirling cycle machine platforms.

The engine 9600 shown in cross-section in FIG. 96A includes generally acrankcase 9610 housing the drive components of the engine and a workspace 9620 containing the working gas and/or fluid and gas and/or fluidcompression and expansion related components. Inside the crankcase 9610is an embodiment of a rocking beam drive mechanism 9601 (the term“rocking beam drive” is used synonymously with the term “rocking beamdrive mechanism”) for an engine, such as a Stirling engine, havinglinearly reciprocating pistons 9602 and 9604 housed within cylinders9606 and 9608, respectively. As discussed previously, rocking beam drive9601 converts linear motions of pistons 9602 and 9604 into the rotarymotion of a crankshaft 9614. Rocking beam drive 9601 has a rocking beam9616, rocker pivot 9618, a first coupling assembly 9619, and a secondcoupling assembly 9621. Pistons 9602 and 9604 are coupled to rockingbeam drive 9601, respectively, via first coupling assembly 9619 andsecond coupling assembly 9621. The rocking beam drive 9601 is coupled toand drives crankshaft 9614 via a connecting rod 9622.

This embodiment shown in FIGS. 96A and 96B is an inverted rocking beamdesign similar to that disclosed in FIG. 92 and incorporates the sameadvantages and benefits as discussed therein. An important advantage ofthe inverted rocking beam arrangement having the crankshaft 9614 locatedrelatively below the rocking beam mechanism 9601 of the machine ensuresthat the structural arrangement and alignment of the piston rods 9624and cross-head coupling means 9634 which connect the piston rods 9624 tothe rocking beam drive 9601 do not have to account for the size of thecrankshaft 9614 and related components. This arrangement facilitates alarger load carrying capacity conrod bearing 9615 on the connecting rod9622, better mechanical advantage developed by the rocking beam 9616 andspace for such larger conrod bearings 9615. The arrangement alsorelieves space constraints of the pistons 9602 and 9604, piston rods9624 and cylinders 9606 and 9608 which can occur with the crankshaftlocated above the rocking beam drive and between the piston shafts 9624.With the rocking beam 9616 now located above the crankshaft 9614, thereare no longer space restrictions around the crankshaft rocking beam 9616and a larger wrist pin bearing 9628 can be provided to better supportthe connecting rod 9622 and rocking beam connection 9601.

Also, with the inverted design, the rocking beam 9616 can be designed toreduce the load on the connecting rod 9622 wrist pin bearing 9628 andconrod bearing 9615 by adjusting the lever arm ratio A and B seen inFIGS. 97A and 97B between the rocking shaft pivot 9718, the connectingrod wrist pin bearing 9728 and between the rocker shaft pivot 9718 andthe lever arm of the piston acting at 9729. For example, as seen in FIG.97A, the bearing load on the connection rod 9728 is greater where theconrod bearing ratio A/B is 1.6 relative to the piston connection. InFIG. 97B, the rocking beam 9616 is shown having a 1.0 ratio whichessentially equates the distances of the two connection points 9728′ and9729 about the rocking shaft pivot 9718 and therefore correspondinglybalances the load on the crankshaft 9614 to be the same or similar tothat developed by the piston shaft 9624 and significantly lower than theload transmitted with the bearing ratio A/B of 1.6. It is to beappreciated that other embodiments of a rocking beam drive besides theinverted rocking beam drive may also incorporate the benefits of thedisclosed rocking beam 9616 as well.

As shown in FIG. 98A, an oil pump 9840 is shown for pumping oil to thebearings in the machine is driven in one embodiment by a helical drivegear (not shown) formed on the crankshaft 9614, and a 90 degree gear setto drive the gear pump 9840. In another embodiment shown in FIG. 98B, aGerotor displacement pumping unit is driven by the crankshaft 9814. TheGerotor pump uses an inner rotor 9844 having one less gear tooth 9846than the surrounding outer rotor 9848. During part of the rotationcycle, the area between the inner and outer rotor increases, creating avacuum that draws fluid through an intake. The area between the rotorsthen decreases, causing compression allowing oil to be pumped out to themechanical parts of the engine. The Gerotor pump is driven coaxially anddirectly from the crankshaft 9814 without the transmission losses of thehelical drive gear, making the engine construction and assembly moreefficient and less expensive than the construction and components of thehelical drive gear to the gear pump. The construction and assembly iseasier because the Gerotor pump is directly driven by the crankshaft9814, whereas there are significant mechanical losses associated withthe previous gear pump. In other embodiments different pumps besides aGerotor pump may be used, which include but are not limited to, gearpumps, piston pumps, rotary gear pumps, hydraulic pumps and diaphragmpumps for example and that other embodiments of a rocking beam drivebesides the inverted rocking beam drive may incorporate the benefits ofthe Gerotor pump or similar direct drive pump.

Returning to FIGS. 96A-E, the alignment of the pistons 9602 and 9604,piston rods 9624 and cylinders 9606 and 9608 in conjunction with thecrankcase 9610 is of critical importance for power transmission throughthe pistons 9602 and 9604 and piston rods 9624, providing for reducedwear on the piston rings and to the dynamic alignment and reciprocatingnature of the piston rods and high pressure piston rod bearings 9630.The crankcase 9610 contains most of the rocking beam drive 9601 and ispositioned below the cylinder housing 9631. The crankcase 9610 defines aspace to permit operation of rocking beam drive 9601 having thecrankshaft 9614 located below rocking beam 9616, a connecting rod 9622,and first and second coupling assemblies 9619 and 9621. Pistons 9602 and9604 reciprocate in respective cylinders 9606 and 9608 as also shown inFIG. 96 and cylinders 9606 and 9608 extend above crankcase 9610, throughthe cylinder housing 9631, and into the heater heads.

The cross-heads 9634 and cross-head bores 9635 have a tolerance that isdifficult to align with that of the mating cylinder liners in thecylinder housing 9631 during what is referred to as “stack up” i.e. thejoining of the separate parts of the vessel, and therefore anydifference in concentricity between these two elements when they areassembled together, creates a potential for misalignment, wherepotentially, the piston rod 9624 could sit askew, or at an angle andtherefore the piston may reciprocate non-coaxially to the cross-heads9634. The cylinder liner bores 9606, 9608, the cylinder gland locatingdiameter and the cross-head locating diameter of the cross-head bore9635 all must be in alignment. To alleviate this issue and potential formisalignment, all three of these diameters are bored together in thesame set-up and essentially simultaneously in the same operationresulting in very close tolerances of the diameters and theconcentricity of these elements is maintained as closely as possiblebased on machining tolerances. These elements may also be manufacturedand bored in other ways as well including but not limited to withalignment jigs and separate boring process that can produce therequisite tolerances to ensure that any angular deviation of the pistonis maintained within an acceptable range.

Also as shown in another embodiment in FIG. 96B, to improve theconcentricity of the piston and piston rod 9624, each piston rod 9624 isprovided with a tapered end 9625 at each end of the rod 9624 to wedgethe first end of the piston rod into the cross-head 9634. The taperedend 9625 facilitates the location, resting and clamping (L,R,C) betweenall elements for a proper location of the piston rod 9624 with thediameter doing the locating, the taper 9625 doing the resting, and a nut9633 of the end doing the clamping. In the wedge connection provided bythe taper, the wedge can lock itself in place because of the loadsdeveloped by the piston. The wedge or tapers on the ends of the pistonrods are essentially jammed more and more firmly into the crossheads9634 at the lower end of the piston rods 9624 and correspondingly intothe piston at the upper end of the piston rod. A nut 9633 may be used tofacilitate the connection with the cross-head 9634 in case the rod comesloose, but in almost every case the wedge will maintain the appropriateconnection of the piston rod 9624 to the piston above, and thecross-head 9634 below.

To facilitate the assembly of the tapered piston rods 9624 where thetaper is essentially a reduction in diameter of the ends of the pistonrod 9624 along a portion of the piston rod, the piston 9602, 9604 ismanufactured from two separate parts, a piston base 9643, and a pistonshell 9645 better shown in FIGS. 96C, D and E. The piston base and shellcan be matingly threaded where the piston base 9643 defines a threadedinner diameter surface wall 9647 corresponding to a threaded outersurface wall 9649 of the piston shell 9645. Other connectionarrangements between the base and shell are possible as well tofacilitate the connection of the two piston elements. The piston base9643 is provided with a receiving bore 9651 which may be a constantdiameter bore, or a tapered bore to receive the tapered end of thepiston rod 9624. To assemble these elements, the tapered piston rod 9624is inserted into the piston base 9643, clamped in place with a desiredpre-load, and then the shell 9645 is threaded onto the base 9643 tocomplete the assembly. The reason for the two-part piston is that toappropriately clamp and pre-load the piston rod 9624 to the base 9643,the assembly process necessitates access to the inside of the piston,and hence the two-part shell and based design facilitates the clampingprocess. Other manufacturing techniques may also be used toappropriately attach the tapered piston rod 9624 and piston 9602, 9604without the necessity for a two-part piston as described above.

Another important aspect of the present embodiment is an increasedvolume of the combustion space in the heater head. To provide morevolume for the combustion of the burner to take place and heat thetubes, an upper most portion 9655 of the cylinders 9606, 9608 isprovided with a larger diameter than the remainder lower portion of thecylinders, giving the cylinders 9606, 9608 to some extent amushroom-shaped profile. The benefit of this includes but is not limitedto the ability to move the heater tubes 9659 farther out from an axialcenter of the cylinders 9606, 9608, thereby increasing the diameter andcombustion volume above the cylinder inside the heater tubes 9659 and/orto accommodate a larger diameter tube to handle more working gas andfluid through the heater tubes 9659.

The heater tubes 9659 which absorb the heat from the burner are criticalto the thermal transfer between the burner and the working fluid insidethe heater tubes 9659. The burner and the flame generated by the burnerextend essentially in the axial center A of the heater tubes 9659 ofeach heater head. Because of this arrangement, the inside surface walls9661 of the heater tubes 9659 are heated more directly and to asubstantially higher degree than the outside wall portions of the heatertubes causing uneven heating of the working gas and/or fluid in theheater tubes 9659. To facilitate a more even heating of the workingfluid and/or gas, an insert 9663 as shown in FIG. 96F is provided insidethe heater tube 9659 which has a star-shaped radial cross-section andmay be, in certain embodiments, twisted in a helical fashion along itslongitudinal axis. Other radial cross-sections could also becontemplated which facilitate channeling of the working fluid/gas sothat contact with varying inner and outer surfaces of the heater tube9659 are accomplished.

The introduction of the insert 9663 into the heater tube defines aseries of substantially separate channels 9665 for directing and evenlyheating the working gas as the working gas passes through the heatertube(s). The channels 9665 in turn define a straight or helical path asthe case may be for the working gas/fluid through the volume of theheater tube 9659 and along the longitudinal axis of the heater tube9659. As the working gas and/or fluid flows through and along thehelical path or channel 9665 in the heater tube 9659, the gas is heatedmore uniformly as it traverses essentially circumferentially around theentire circumference of the heater tube 9659 between the cooler backwall and the hotter inside surface wall 9661. Additionally, the insert9663 provides a more direct heat transfer of the heat provided by theburner from the inside surface wall 9661 to the back wall of the heatertube 9659, and also around the entire circumference of the heater tube9659 where the insert 9663 itself is a heat path or conduit foruniformly heating the entire circumference and length of the heater tube9659.

Another benefit of using such an insert 9663 is the increased flowvelocities of the working gas through the heater tubes 9659 which inturn leads to a better heat transfer to the working gas. The rate ofconductive heat transfer can be understood as Q is equal to mass flowrate times the change in temperature T, so for any given temperaturechange, as the velocity increases through the smaller channels 9665defined by the insert 9663 so does the heat transfer to the working gas.

The present embodiment of the Stirling cycle engine maintains theworking space 9620 and the working gas and/or fluid at a relatively highpressure, generally in the range of 1200-1800 psi, and more preferablyabout 1500 psi. It is of course necessary to ensure that the working gasand/or fluid is essentially sealed in the working space 9620 so that itdoes not escape into the crankcase 9610 and the environment. A criticalplace for such leakage of working fluid to occur is around the pistonrods 9624, which extend and reciprocate between the working space 9620and the crankcase 9610. To minimize such leakage, a high pressure pistonrod seal 9630 is provided below the respective cylinders 9606 and 9608and between the working space 9620 and the crankcase 9610.

High Pressure Rod Seals

With a significantly higher pressure in the working space 9620 relativeto the crankcase 9610, a certain amount of working gas is anticipated toleak through the high pressure rod seals 9630. However, it is imperativeto minimize the leakage without significantly affecting thereciprocating efficiency of the pistons and the engine. Also, as will bediscussed in further detail below, an airlock and working fluidrecapture system may be used in conjunction with the high pressure sealsto capture certain amounts of such leaking working gas and/or fluid. Anyworking gas which leaks into the air lock between the working space 9620and the crankcase 9610 can be drawn into an accumulator and suppliedback into the workspace when necessary. Before more completelydiscussing such an airlock and recapture of working fluid, the presentdiscussion is focused on the use of the high pressure rod seals 9630between the working space 9620 and the crankcase 9610 to ensure the mosteffective working fluid pressure and gas containment.

A mechanical embodiment of the high pressure rod seal 9930 is shown in atest rig assembly in FIG. 99A for purposes of clarity. It should beunderstood that such a rod seal is intended to be utilized not only inthe Stirling engine embodiments described herein but also in otherengines or mechanisms with similar reciprocating pistons. A rod sealcavity 9932 is defined in the test rig for housing the rod seal 9930.The high pressure wedge rod seal 9930 in FIG. 99A includes severalcritical components: a rod sleeve 9940, a rod lower seal support 9942and a clamp spring 9950. The rod sleeve 9940 defines a passage throughwhich the piston rod 9924 is inserted and frictionally slidably engages.The sleeve 9940 has an inner diameter sized similar as the outerdiameter of the piston rod 9924 and is made of a low friction materialso that the piston 9924 slides easily through the passage as itreciprocates. The rod sleeve 9940 is provided at one end with a wedgeportion 9946, which is shaped to fit into and be springably maintainedagainst a mating seat defined in the rod lower seal support 9942. Thenature of the wedge portion 9946 of the sleeve 9940 in conjunction withthe lower seal support 9942 aids in the self-alignment of the seal, evenwith the reciprocating rod 9924 in motion, so that initial manufacturingtolerances of the sleeve 9940 and lower seal support 9942 as a whole areless strict and a tight seal between the sleeve 9940 and piston rod 9924is maintained over time. The lower seal support 9942 can include aseries of fittings such as gaskets, washers or o-ring seals 9944, whichseal the base against the supporting surfaces of the rod seal cavity9932 to ensure neither working gas or pressure passes or escapes from,what in a working Sterling engine is defined as, the work space.

The sleeve 9940, and hence the wedge 9946, maintains its tight fittingcontact with the mating seat in the lower seal support 9942 under theinfluence of a load applied by the clamp spring 9950 shown in FIG. 99A.The clamp spring 9950 influences the rod sleeve 9940 against the lowerseal support 9942. The clamp spring 9950 as shown in FIG. 99A iscomprised of an upper collar 9955 and a lower collar 9956 sandwiching aplurality of coil springs 9957 therebetween. The lower collar 9956 has acircumferential chamfered inner surface 9958, which bears on a matingsurface on the wedge portion 9946 of the rod sleeve 9940. Based on apreload applied to the clamp spring 9950, the lower collar 9956 thusinfluences the rod sleeve 9940 into sealing contact with the lower sealsupport 9942 to ensure that significant leakage of working gas andpressure cannot escape from the working space 9620.

The clamp spring 9950 is designed to take up any play which occurs inthe seal 9930 due to wear. The wedge portion 9946 of the sleeve 9940will wear over time against the seat in the lower seal support 9942 andas these contact surfaces wear the clamp spring 9950 essentially extendsdue to the spring bias and continues to provide axial and radial forcesagainst the wedge portion 9946. Clearance spaces are provided betweenthe lower collar 9956 and the lower seal support 9942, so that as thewedge portion 9946 and lower seal support 9942 wear, the clamp spring9950 can expand and still maintain appropriate axial and radial forceson the sleeve 9940.

In another mechanical embodiment of the high pressure piston rod seal9930′, shown in better detail in FIG. 99B, a substantially symmetricalhemispherical shaped piston sleeve 9960 is supported by an upper sealsupport 9965 and a lower seal support 9966 inside a seal cavity definedinside a seal housing 9951. The symmetry of this hemispherical shapedpiston sleeve 9960 provides more consistent wear across the length ofthe sleeve 9960 as compared to the wedge rod seal 9930 described abovewhich focuses the radial wear at one end of the sleeve. Thehemispherical surface 9963 of the piston sleeve 9960 bears on an innerrespective bearing surface of each of the upper and lower seal supports9965, 9966. A wear support clamp 9967 is provided axially disposed abovethe upper seal support 9965 which forces the upper and lower sealsupports 9965, 9966 into biased contact with the piston sleeve 9960. Agap G may be provided between the upper and lower seal supports 9965,9966 to accommodate any wear that may occur on abutting surfaces in theseal. As wear occurs, the abutting surfaces in the seal may be reducedso that as the sleeve bearing wears, the upper and lower seal supports9965, 9966 are biased towards one another by the support clamp 9967. Thegap G permits the upper and lower seal supports 9965, 9966 to movecloser to one another as the seal wears without interfering with oneanother and so maintaining contact with the hemispherical shaped outersurface 9963 of the piston sleeve 9960.

A still further embodiment of a high pressure rod seal shown in FIG.100A includes a spring energized lip seal 10003 generally comprising aseal jacket, made from PTFE or graphite for example, and a spring (notshown) circumferentially secured within a groove or between lips 10007of the seal 10003. When the spring energized lip seal 100003 is seatedin the housing, the spring lip seal 10003 is under compression, forcingthe jacket lips 10007 against the respective adjacent walls of the sealblock 10011 and the surface of the reciprocating piston 10024, therebycreating a leak free seal. The lip seal 10003 provides permanentresilience to the seal jacket 10005 and compensates for jacket wear andhardware misalignment or eccentricity. System pressure also assists inenergizing the seal jacket 10005. Spring loading assisted by systempressure provides effective sealing at both high and low pressures.Spring energized lip seals are highly durable and designed for static,rotary and reciprocating applications in temperatures from cryogenic to+600 F as well as pressures from vacuum to 25,000 psi, and to survivemost corrosive environments.

A spring cup retaining cylinder 10008 is set around the piston rod 10024and supported on a lower collar 10006. The retaining cylinder 10008maintains a circumferential space about the piston rod 10024 in whichthe lip seal 10003 is maintained. The lip seal 10003 can be a PTFE andgraphite ring supported around an outer circumference by the retainingcylinder 10008 and frictionally slidably engages the piston rod 10024 tocreate the high pressure spring energized lip seal. The spring (notshown) inside the lip seal 10003, along with the higher pressure of theworking space, forces the lip seal 10003 against the respective pistonrod 10024 and retaining cylinder wall, and also maintains the lip seal10003 set down in the retaining cylinder 10008 generally against thelower collar 10006.

A hydraulic embodiment of a high pressure piston rod seal can facilitatean efficient and long term seal between the working space and theairlock. FIG. 100B discloses a hydraulic high pressure piston rod seal10021 set inside the rod seal cavity of a test rig. A rod seal sleeve10023 circumferentially encompasses the piston rod 10024 and defines apressure space 10025 between a wall of the test rig and an outer surfaceof the rod seal sleeve 10023. A hydraulic fluid pressure line 10027communicates with pressure space 10025 to provide the appropriate fluidpressure to maintain the rod seal sleeve 10023 in sealing engagementwith the piston rod 10024. A sensor (not shown), such as apiezo-electric pressure sensor, can be provided in the pressure space10025 and on the rod seal sleeve 10023 to ensure that the appropriatepressure and flexure is actuating the rod seal sleeve 10023 andproviding the appropriate sealing pressure against the piston rod 10024.The inner surface of the rod seal sleeve 10023 slidably engages alongthe piston rod 10024 as the rod reciprocates and the rod seal sleeve10023 is influenced radially inwards by the hydraulic pressure fluid inthe pressure space 10025. As the rod seal sleeve 10023 wears, thehydraulic fluid pressure in the pressure space 10025 can be increased toensure that the rod seal sleeve 10023 is motivated radially towards thepiston rod 10024 to maintain slidable engagement with the piston rod.

It is to be appreciated that the above disclosed embodiment of highpressure rod seals are intended only as examples and that the machinesdescribed herein are not limited to these examples, and that otherembodiments of high pressure rod seals may also be used to ensure thatthe high pressures used in Stirling engines, or any other engine forthat matter, are maintained in the appropriate working space, crankcaseand other engine compartments as necessary.

Rolling Diaphragm Seal

Turning to FIGS. 101A and 101B, and referring back also to FIGS. 13 A-G,in certain embodiments of the present invention a rolling diaphragm10190 is used in conjunction with the piston rods 10124 to prevent theescape of lubricating fluid from the crankcase 10110 up past the rods10124 and into the working space 10120 and regenerator. If thelubricating fluid necessary for the rocking drive can bypass the pistonrod seals, it can potentially damage the working space, clog theregenerator and contaminate the working fluid or gas of the engine inthe cylinders.

To facilitate the appropriate rolling and flexing of the diaphragm10190, a pressure differential is maintained across the rollingdiaphragm 10190 so that preferably the pressure above the diaphragm10190 is slightly greater than the pressure in the crankcase. The sealis thus essentially inflated into the crankcase, which facilitates thediaphragm 10190 maintaining its desired form as it rolls and flexes withthe reciprocating piston rod 10124. This alleviates stresses on thecircumferential sealing points so the seal is not compromised. It isgenerally necessary to place a differential of approximately 15 PSIacross the diaphragm 10124 to properly inflate the seal so that itconforms to the shape of the bottom seal piston 10195 as it moves withthe piston rod 10124. It is to be appreciated that the pressuredifferential maintained across the rolling diaphragm 10190 is notlimited to 15 PSI. Rolling diaphragms made of stronger materials orhaving a particular shape may be able to sustain a higher differentialor operate at a lower differential as the case may be. In embodiments ofthe stirling cycle engine where the working space 10120 is at arelatively high pressure 1500 PSI-1800 PSI, the crankcase 10110 must becharged with a pressure for instance of 1485 PSI, which is approximately10-15 PSI less than that of the working space at 1500 PSI. Although itis possible to regulate these larger pressures to maintain the 10-15 PSIdifference across the diaphragm, it is difficult and adds to thecomplexity of the machine.

The rolling diaphragm 10190 may be manufactured by injection molding orhot compression molding. In hot compression molding of the rollingdiaphragm 10190, it can be more difficult to control material propertiesbut injection molded diaphragms have shown in testing a bettertransition of dynamic stresses across the profile of the rollingdiaphragm 10190 as it transitions and rolls with the reciprocation ofthe piston rod 10124. Testing on the materials used to fabricate therolling diaphragm 10190 indicate chopped fiber is most successful forexample but not limited to, nitrile with Kevlar fiber or Fab-Air®.

FIGS. 101A and 101B disclose an embodiment of the rolling seal ordiaphragm 10190 having a profile which facilitates the dynamic rollingtranslation of the diaphragm. As discussed in previously herein, andincorporated herein by reference in its entirety in the presentdiscussion, the pressure differential that is placed across the sealallows the seal to act dynamically to ensure that the rolling diaphragm10190 maintains its form throughout its dynamic range of motion. Aspreviously discussed, the pressure differential causes the rollingdiaphragm 10190 to conform to the shape of the bottom seal piston 1310with reference to FIG. 13A as it moves with the piston rod 10124, andprevents separation of the diaphragm 10190 from the surface of thepiston rod 10324 during operation. It is desirable to lower the amountof inflation of the rolling diaphragm 10190 without the diaphragmbuckling or separating, i.e., deviating from a consistent dynamic axialand radial rolling of the diaphragm 10190 along the diaphragm profilewith the axial reciprocation of the piston rod 10124. As discussedabove, the inflation of the diaphragm is provided by the pressuredifferential across the rolling diaphragm 10190. To accomplish this, ithas been found that particular structural profiles facilitate theconservation of material and consistent rolling of the diaphragm.

The cross-section in FIG. 101A-B shows a profile view of the molded formof the diaphragm of the present embodiment about a diaphragm axis L. Forpurposes of describing the diaphragm structure the inner edge 10192 asbeing the top 10194 of the diaphragm and the outer edge 10193 is thebottom of the diaphragm as shown in the figures. The diaphragm has alateral wall 10190 extending axially and radially relative to axis Lfrom the inner edge 10192 to the outer edge 10193; the lateral wall iscomposed of several sections. A top fillet section 10198 turns thematerial approximately 90 degrees from the top of the diaphragm 10190 asshown, to a sidewall section 10196 substantially parallel to the pistonrod 10124 and axis L. The sidewall section 10196 in turn then turnstowards the outer edge 10193. Before reaching the outer edge 10193, thesidewall section merges contiguously into a chamfer section 10199, whichwhile still depending axially from the sidewall section 10196, extendsfrom the sidewall 10190 in a greater radial degree relative to axis L toconnect with the outer edge of the diaphragm 10193. The sidewall section10196 may be parallel to the axis L or may also have a radial componentwhich slightly angles the sidewall section 10196 radially away from theaxis L. In either case the chamfer section 10199 extends to a greaterradial degree from axis L than the sidewall section 10196. A bottomfillet 10197 connects to the outer edge 10193 defining the bottom of thediaphragm as drawn. The outer edge 10193 like the inner edge 10192 isprovided with a thickened circumferential lip, which can be securedinside a matching groove formed in the vessel joint.

The cross-section shown in FIGS. 102A and 102B is a profile view of themolded form of another embodiment of the rolling diaphragm 10290 of thepresent embodiment about a diaphragm axis L. Like reference numbers forthis embodiment correspond to the same or similar elements in theprevious rolling diaphragm embodiment. For purposes of describing thediaphragm structure, the inner edge 10292 is the top of the diaphragmand the outer edge 10293 is the bottom of the diaphragm 10290 as shownin the figures. The diaphragm 10290 has a lateral wall 10296 extendingaxially and radially relative to axis L from the inner edge 10292 to theouter edge 10293; the lateral wall here is again composed of severalsections. A top fillet 10294 section turns the material approximately 90degrees from the top of the diaphragm 10290 as shown, to the sidewallsection 10296, which depends both axially and radially outwards towardsthe bottom of the diaphragm along the axis L. Before reaching the outeredge 10293, the sidewall section 10296 merges contiguously into a bottomfillet 10299 to extend towards the outer edge 10293 of the bottom of thediaphragm as drawn. An outer lip 10197 similar to the thickenedcircumferential lip 10295 of the inner edge 10192 is provided, which aresecured inside a matching groove formed in the vessel or crankcase jointwhich secures and seals the outer edge 10293 of the diaphragm.

The injection molding of the diaphragm is important because the gatingmethods and other molding techniques, characteristics, methods andspecifications can affect the fiber alignment and molecular alignment ofthe diaphragm material during the molding process. These materialcharacteristics are important because this can affect the hoop stress ofthe diaphragm. For example, if the material is gated at one end andoverruns an opposing end of the mold, the fibers can be aligned in aparticular direction to optimize the hoop strength of the diaphragmwhile potentially enhancing the flexible and rolling characteristics ofthe final diaphragm element.

It is very important in the dynamic rolling actuation of the diaphragms10190, 10290 that no imperfections or particles including fluidparticles such as oil droplets are disposed on the surfaces of thebottom seal piston or on the adjacent vessel wall surrounding the bottomseal piston. Such fluid particles, most likely oil, are detrimental tothe rolling actuation of the diaphragm against the respective crankcasesurfaces, because they cause stress points on the diaphragm.

Turning to FIG. 103 another embodiment of the rolling diaphragm includesa first and second rolling diaphragm 10391, 10393 to make what isessentially a double bellows system 10392. A double bellows system 10392can facilitate the elimination of the 10-15 PSI pressure differentialbetween working space and airlock and/or crankcase by providing theappropriate expansion pressure between the double bellows themselves.The double bellows include first and second rolling diaphragms 10391,10393 which are oppositely and axially aligned along the piston rod, anddefine a space therebetween with a light oil contained between thediaphragms and pressure charged between the double bellows. Theincompressible oil prestresses the diaphragms and facilitates theconsistent rolling of the diaphragm as the piston rod 10324 reciprocatesalong its axis.

Airlock and Working Fluid Recapture System

The power, life and value of a Stirling engine can be maximized bybuilding an oil lubricated drive and sealing the work-space from theoil. Oil lubricated drives allow high powers and are inexpensivecompared to drives based on rolling elements. It is essential to isolatethe oil in the drive from the workspace. Even oil mist will migrate tothe hot end of the working space, where the oil will breakdown and theresulting carbon will clog the heat exchanger. Flexible membranes orbellows such as the rolling diaphragms discussed above that attach tothe moving piston rod and the structure provide an oil and gas tightseal between the oil filled crankcase and the workspace, ensuring thatthe lubricant is maintained in the crankcase. In order to function forthousands and millions of cycles, a small pressure difference must bemaintained across the bellows.

An important aspect of the rolling diaphragm and oil lubricatedcrankcase relates to the use of an airlock 10401 and an airlock pressureregulation system 10411 as shown in FIGS. 104A and 104B. The airlockpressure regulation system 10411 provides the benefit of ensuringworking gas escaping from the working space 10403 is returned to theworking space, provided that the working gas does not leak into theenvironment or atmosphere, which would require replenishment of theworking gas, and that an appropriate pressure differential is maintainedacross the rolling diaphragms as described above. The airlock pressureregulation system 10411 permits an easily serviceable bottom end i.e.crankcase 10410 if, as in the embodiment disclosed in FIG. 105, thecrankcase is intended to be maintained essentially at atmosphericpressure.

As shown in FIG. 104A relating to a pressurized crankcase 10410 atapproximately 1485 PSI, in order to maintain an appropriate workingspace pressure and airlock pressure regulation, an airlock space 10401is provided between the working space 10403 and the crankcase 10410 at apressure of, for example 1500 PSI, so that the substantially greaterpressures in the working space 10403 should not significantly influencethe air lock space 10401 and any pressure and working gas leaking fromthe working space 10403 into the air lock can be captured andaccumulated as described below with respect to the airlock pressureregulator and returned to the airlock and working space and not merelyescape into the crankcase and environment.

It is to be understood that airlock space 10401 is intended to maintaina constant volume and pressure necessary to create the pressuredifferential necessary for the function of rolling diaphragm 10490 aspreviously described. In the present embodiment the airlock 10401 may ormay not be sealed off from working space 10403 with high pressure rodseals 10430. In any case, the pressure of airlock space is desired to bemaintained at essentially 1500 PSI and equal to the mean pressure ofworking space 10403. The pressure in the working space 10403 may vary atleast +/−300 PSI so the intention of the airlock space 10401 is toinsulate the diaphragms from such fluctuations and maintain itself ataround the necessary pressure, by way of example here 1500 PSI, relativeto the 1485 PSI charged in the crankcase 10410. To facilitate theequalization of pressures between the working space 10403 and theairlock space 10401, a small opening or pressure equalization orifice10404 communicates between the working space 10403 and the airlock space10401. The crankcase 10410 must be charged to 1485 PSI, and bemaintained at approximately 15 PSI less than the airlock space 10401 sothat the appropriate pressure is applied to the rolling diaphragm 10490to ensure the proper dynamic movement of the diaphragm.

In this pressurized crankcase 10410 embodiment an airlock pressureregulator 10411, a pump 10412 and relief valve system is providedbetween the crankcase 10410 and the air lock space 10401 to maintain theexemplary 15 PSI pressure differential therebetween. Other predeterminedpressure differentials may also be maintained depending on the diaphragmmaterial and the design of the entire airlock pressure regulationsystem. In its most general form, an uptake line 10416 communicates fromthe pressurized crankcase 10410 to a filter 10418, a pump 10412 (havinga check valve on its outlet), and a pressure regulator 10413 in parallelwith the pump 10412 and filter 10418 for returning pressurized workinggas back to the air lock 10401 and so maintains the pressuredifferential between the airlock space 10401 and the crankcase 10410 andconsequently across the rolling diaphragm 10490. This airlock pressureregulator system 10411 is described more completely with respect to FIG.104B.

The airlock pressure regulator 10411 regulates the pressure differencebetween the airlock 10401 and the crankcase 10410. When the engine isturning, the airlock pressure regulator 10411 keeps the airlock pressurepreferably 10 to 14 PSI above the crankcase pressure although a range of5 to 20 PSI is possible and other pressure differentials can beaccomplished by the regulator as well. When the engine is off, theairlock pressure regulator 10411 keeps the airlock pressure preferablyless than 15 PSI above the crankcase pressure and not more than 5 PSIbelow crankcase pressure. It is permissible to have a greaterfluctuation of pressure differential when the engine is off since thereis little or no dynamic forces being applied to the rolling diaphragms10490 via moving pistons.

The airlock pressure regulator 10411 performs several importantfunctions. The airlock pressure regulator 10411 uses a pump 10412 tomove pressurized gas from the lower pressure crankcase 10410 into theairlock 10401, thereby maintaining the airlock 10401 at a higherpressure. The airlock pressure regulator 10411 relieves excess pressurebetween the airlock 10401 and crankcase 10410 volumes. A bidirectionalregulator 10413 vents some of the airlock gas into the crankcase 10410,when the airlock pressure is preferably 15 PSI above the crankcase andvents in the opposite direction, venting gas from the crankcase 10410 toairlock 10401, when the airlock pressure is more than 5 PSI below thecrankcase pressure. Also, a filter 10418 in the airlock pressureregulator 10411 filters out the oil from the crankcase gas before itenters the airlock volume.

The components of the preferred embodiment are the mechanical pump10412, the bidirectional pressure regulator 10413, an oil filter 10418,a pump pressure switch 10417 to control the pump 10412 and a controllerpressure switch 10419 to signal the engine controller C. An example ofthe mechanical pump is the Linear AC 0410A pump by Medo. Other pumpscould certainly be used as well. The important qualities of the pump arethe ability to operate in a high pressure inert environment, long life,no maintenance and quiet. Solberg Mfg. produces a line of oil-misteliminators, i.e. filters, that are compact, effective and can holdenough oil for several thousand hours of operation. In a preferredembodiment, the bidirectional regulator 10413 allows pressure flow whenthe design pressure difference has been exceeded in either direction.Pump pressure switch 10417 operates the pump when the pressuredifference between the airlock 10401 and the crankcase 10410 ispreferably less than 10 PSI for example. Pump pressure switch 10417includes a predetermined dead band, or range, that keeps the pump 10412on until the airlock pressure is for example 14 PSI above the crankcasepressure. Controller pressure switch 10419 signals to the controller Cthat the airlock pressure is at least 5 PSI, for example, above thecrankcase pressure. This insures that the engine will not turn until theairlock pressure is sufficiently greater than the crankcase pressure.The rolling diaphragms 10490 could tear if moved without such pressuredifference across them. A fill source 10414 may be connected with theairlock to replenish the pressurized vessel charging and workinggas/fluid if necessary.

FIG. 104C is a specific embodiment of the bidirectional regulator 10413showing the pump 10412, oil filter 10418 and a spool valve 10441 whichoperates between an airlock port 10449, a crankcase port 10451 and apump port 10453 according to the pressure differentials between thecrankcase pressure and airlock pressure. In this case, alternative tothe pressure switches 10417, 10419 described above a proximity sensor10425 for determining location of the spool 10441 via a target magnet10426 is used to control the pump 10412 and if necessary to signal theengine controller C. The spool valve 10441 is biased by a primary spring10443 against the airlock over-pressure and an underpressure reliefvalve 10445 is biased by an inner spool spring 10447. Observing FIG.104D-104H the spool is shown in certain positions: in (1) is shown thespool influenced open by the spring where the airlock pressure is low sothat the airlock port 10449 now communicates to pump port 10453 toreceive pressurized gas from the pump 10412, in (2) the spool 10441 isshown where the airlock pressure is within normal limits so the airlockport 10449 is closed by spool 10441 and the spool is still displacedenough according to the proximity sensor 10425 to cause operation of thepump 10412, even without flow from the pump to the airlock. In (3) thespool 10441 is shown where the airlock pressure is again within normallimits so the airlock port 10449 is closed by spool 10441 and the spoolis now displaced so that the proximity sensor 10425 does not turn on thepump. Either one or two proximity sensors 10425 are shown in FIGS.104C-D, however any desired number and type of proximity sensors may beused in normal operation in other embodiments. In (4) the spool 10441 isshown with the airlock pressure is high so that the airlock port 10449is connected to crankcase port 10451 and pump is disabled while airlockpressure is reduced. (5) is a case where the engine is shut down sothere is no power to the pump and the airlock pressure is extremely lowand to keep the diaphragms from being damaged, the airlock port 10449 isconnected directly to the crankcase through the underpressure reliefvalve 10445 which opens to provide direct pressure relief through thespool 10441 so that the crankcase pressure and airlock pressure are atleast equalized.

In another embodiment of the pressure regulator 10401 shown in FIG.1041, the bidirectional regulator 10413 is replaced with a back pressureregulator 10431 which provides for one way pressure flow from theairlock 10401 into the crankcase 10410 should the pressure differentialexceed for instance 15 PSI. To accommodate flow in the other directionfrom the crankcase to the airlock, a check valve, or pair of checkvalves 10433, 10435 can be provided in a separate path. This ensuresthat the crankcase will not be pressurized higher than the airlock. InFIG. 105, the crankcase 10510 is intended to be maintained atatmospheric pressure. This is a critical improvement of the presentinvention as it provides for a more easily serviceable lower unit on thevessel without the need to recharge the crankcase 10510 should work needto be done inside the crankcase 10510 and also provides that asignificantly lighter crankcase housing is necessary to contain thedrive components. In this embodiment of the airlock pressure regulatorsystem 10511 the airlock space 10501 is maintained essentially atatmospheric plus 15 PSI and therefore any pressurized working gas whichescapes from the working space 10503 into the airlock 10501 needs to beremoved from the airlock 10501 and returned to the working space 10503.To accomplish this, in its simplest form a first relief valve 10520means is provided in an uptake line 10522 communicating with the airlockspace 10501 so that any pressure greater than 15 PSI is relieved fromthe airlock 10501 and passed via a pump 10512 into an accumulator 10523outside the working space 10503, airlock space 10501 and crankcase10510. From the accumulator 10523 a return line 10525 includes a secondrelief valve 10521 means which opens to permit recharging of the workingspace 10503 with pressurized gas from the accumulator 10523 should thepressurized working gas in the working space 10503 fall below 1500 PSI.It is to be appreciated that the balancing of this pressurized systemmay include other pressure considerations across the first and secondrelief valves 10520, 10521, particularly with respect to the variationwhich can occur in the working space 10503 where the pressure can swingplus or minus 300 PSI during the Stirling cycle itself.

When the engine is running, a mechanical pump 10612 defined by a cavity10608 in the piston rod may be utilized to reduce the load and work doneby the above described airlock pressure regulator system. As seen inFIG. 106, the mechanical pump 10612 of the piston rod 10624 is added tothe airlock pressure regulation system to reduce the load on theelectrical system during operation. A check valve 10605 receivescrankcase pressure through an intermediate passage 10607 from thecrankcase. The check valve 10605 opens when the airlock pressure hasdropped too low relative to the crankcase pressure and pressurized gasfrom the crank case is drawn into the piston cavity 10608 as the pistonrod 10624 reciprocates. The piston rod cavity 10608 is defined by areduced diameter portion of the piston rod which essentially defines themechanical pump 10612 itself. As the piston rod 10624 reciprocates thepiston rod cavity 10608 is reduced in size as shown by the right-handpiston, pumping the pressurized gas into the airlock space 10609. Inthis way during engine operation the airlock can be efficientlyreplenished with sufficient pressurized gas should its pressure drop toolow. An outlet check valve 10611 is provided between the airlock and thecrankcase so that pressure in the airlock which exceeds the desireddifferential can be reduced from the airlock space 10609 into thecrankcase. The mechanical pump 10612 defined by movement of the pistonrod 10624 does not operate at engine startup because there is nomechanical operation of the engine, however the airlock pressureregulator system must be operational during startup operations.

Cooler Liner Diameter Reduction

As explained previously with respect to FIGS. 65D-E, the heater tubescommunicate with a heat exchanger which circumferentially surrounds eachcylinder. The heat exchanger of the present embodiment described inFIGS. 18A and 18B provides cooling for the working gas/fluid during theappropriate portion of the stirling cycle. The heat exchanger 10705 issupplied with cooling water through coolant tubing which communicateswith a heat sink such as the environment via a radiator (not shown).Generally, the coolant water picks up heat through the heat exchanger inthe vessel from the hot working gas, and the coolant water then ispumped to the radiator where the heat is transferred to the environment.

The heat exchanger 10705 shown in FIG. 107A surrounding each respectivecylinder is provided with a water jacket sleeve 10704 having an innersurface defining a channel to allow passage of the cooling water throughan interfacial area 10706 between the inner surface of the water jacketsleeve 10704 and an outer surface of a cooler liner 10702. The coolerliner 10702 also has an inner surface 10708 which directs the flow ofhot working gas along the inner surface to facilitate the transfer ofthe heat through the cooler liner 10702 to the coolant water in theinterfacial area. A goal of the described structure is to increase theheat transfer surfaces within the interfacial area 10706 for absorbingheat from the hot working gas and the heat exchanger so as to improveheat transfer between the working gas and the coolant water.

The water jacket sleeve 10704 surrounds the cooler liner 10702 and formsthe heat exchanger 10705 which cools the working fluid during theappropriate portion of the Stirling cycle. The cooler liner 10702directs the flow of the working gas along the inner surface of thecooler liner 10702. An improvement of the presently described engine isan increased heat transfer surface area in the heat exchanger where theouter diameter of the cooler liner 10702 is reduced to increase theinterfacial area 10706 with a plurality of extended surfaces, forinstance, longitudinal arranged fins 10707, or pins, provided around theoutside diameter of the cooler liner 10702 and extending into theinterfacial area 10706 between the inner surface of the water jacketsleeve 10704 to increase the surface area of the heat exchangingsurfaces and provide more efficient cooling of the working gas/fluid.The outer diameter of the cooler liner wall 10708 can be reduced to anextent so that the cooler liner wall 10708 is relatively thin, ascompared to the radial length of the longitudinal fins 10707, or pins inthe interfacial region 10706 between the cooler liner 10702 and theinner surface of the heat exchanger 10705. The inner wall 10708 of thecooler liner is generally maintained an appropriate diameter toaccommodate the working gas flow from the heater head and cylinder. Theinner diameter of the liner is provided with axially arranged fins 10707to direct the flow of gas along the inner wall of the liner andfacilitate the transfer of heat out of the working gas, through thecooler liner and into the coolant water. (Drawing of cross-section ofcooler liner and water jacket sleeve??)

It is also important to ensure that the stationary seals utilized in theheat exchanger are to the extent possible redundant and not compromised,particularly where the water jacket sleeve 10704 and cooler liner 10702should sufficiently maintain the coolant water in the interfacial region10706 between these elements and the working fluid inside the coolantline 10702 r. As shown in FIG. 107B, the heat exchanger 10705 in thepresent embodiment has an outer surface which abuts the inner surface ofthe vessel and is sealed with respect to the vessel by an upperstationary seal 10710 and a lower stationary seal 10711. Similarly thecooler liner 10702 inside the heat exchanger 10705 is sealed withrespect to the inner surface of the heat exchanger by an upper seal10713 and a lower seal 10715. A top surface of each of the cooler liner10702 and the heat exchanger 10705 are formed and both support the baseof the heater heads 10703 and provide the communicating interface forthe working gas between the heater tubes 10709 and the heat exchanger10705. An additional or redundant seal can be added at the intersectionbetween the cooler liner 10702 and the heat exchanger 10705 adjacent thetop surface of each element which supports the heater heads 10703. Thisredundant seal 10712, for example a 45 degree o-ring, is located axiallyspaced above the upper stationary cooler liner sealer 10710 and extendscircumferentially around the entire joint between the heat exchanger10705 and the cooler liner 10702. The addition of the heater head baseas it is supported on the top surfaces of the liner 10702 and heatexchanger 10705 compresses the redundant seal into the joint and addsredundancy to the system to prevent the escape of cooling water and/orworking gas/fluid from the working space.

In a further improvement to the drive system of the present invention amore easily constructed and easy to maintain connection between the linkrod 10826 and the rocking beam 10816 is described. Referring now to FIG.108A, a rocking beam drive mechanism 10801 is shown. In this embodiment,the rocking beam drive mechanism has pistons 10802 and 10804 coupled totwo rocking beam drives 10801. In the exemplary embodiment shown moreclearly in FIG. 108B, the link rod 10826 is coupled at a first end tothe piston rod via a link rod upper pin 10832, and a second end of thelink rod 10826 may be coupled to one end of a link rod lower pin 10832attached to the yolk of the rocking beam 10816. The link rod lower pin10832 had been previously accomplished by press fitting a pin 10823 intoa passage of the link rod 10826, and with bearings provided on eitherside of the link rod 10826 and around the pin 10823, the second end ofthe link rod is secured to the rocking beam drive 10801 in a yolk 10825.The pin 10823 extends into respective pin passages in the yolk 10825 ofthe rocking beam 10816 in order to complete the link rod lower pin 10832structure. A bearing is also provided between the pin passages in therocking beam 10816 and the pin 10823 to facilitate the pivoting of thelink rod and pin relative to the rocking beam 10816.

The present embodiment eliminates the need for a press fit of the pininto the passage in the second end of the link rod 10825. The press fitmade it difficult to maintain, fix assemble and disassemble in anymanner this end pivot structure during maintenance of the engine. Asseen in FIG. 108B the link rod lower pin 10823 is provided to beinserted with a loose fit into and through the passage in the second endof the link rod 10826. A bearing 10822 may be provided around the linkrod lower pin 10823 on either side of the link rod 10826, and the widthof the bearing 10822 is reduced in order to fit a retaining ring 10828onto the pin 10823 adjacent each bearing to retain the pin 10823 andbearings axially aligned in the passage of the link rod 10826 and in theyolk 10826. With the pin 10823 and bearings essentially axially fixed bythe retaining rings 10828 to the link rod 10826, the pin 10823 and linkrod passage can have a loose fit so that the pin 10823 can be easilyremoved from the link rod 10826, when disassembly is necessary, merelyby removing the retaining rings 10828 and sliding the pin 10823 out ofthe link rod passage. A lubricating oil passage 10829 may be provided inthe link rod lower pin 10823 to communicate with a oil passage 10838 inthe link rod 10826 and provide oil to the bearings 10822 and therespective surface of these pivoting components.

The link rod upper pin 10832 is similarly arranged with a loose fit withthe first or upper end of the link rod 10826. A bearing 10834 in thiscase is provided directly between the bearing surface of the upper pin10832 and an inner surface of the upper link rod passage. A pair ofretaining rings 10836 are applied to grooves in the ends of the upperpin 10832 to maintain the pin in its axial placement in the cross head10840. The bearing 10834 and respective bearing surfaces can be suppliedwith lubricating oil via the oil passage 10838 in the link rod 10826

The arrangement of the Stirling machine discussed above is generallyreferred to and shown as having a vertical orientation, i.e. with thepistons reciprocating generally perpendicularly aligned relative to ahorizontal support surface or ground surface. In another embodiment ofthe present Stirling cycle engine 10903 shown in FIGS. 109A and 109B theengine may be horizontally arranged, i.e. with the pistons 10905, pistonrods 10907, heater heads 10911, cross heads 10913 etc., being arrangedand reciprocating in a horizontal orientation relative to a groundsupport surface as opposed to the vertical orientation discussed above.One of the significant challenges in such a design is the arrangementand structure of the oil cooling system in the crankcase 10915 where itimperative to ensure that the mechanical elements of such a horizontalcrankcase such as the cross heads 10913, rocking beam 10919 and othercrankcase components and drive elements are sufficiently supplied with afree flow of oil through the crankcase and back to the oil sump andpump.

As seen in FIG. 109B and by way of general example, the oil coolingsystem comprises a central oil supply line 10921 disseminating a flow ofoil directly to each of the cross head bores 10923 through radial oilpassages 10925. Oil drains down by gravity in the crankcase 10915 intooil sump 10931 which can then be re-circulated back to the central oilsupply line 10921 via a pump 10935 through main line 10937 whichcommunicates eventually with central oil supply line 10921. It is to beappreciated that other oil supply arrangements and orientations can alsobe accomplished, and that the embodiment described with respect to FIGS.109A-B and the horizontal arrangement of the engine and crankcasecomponents is merely exemplary with respect to these figures.

In another embodiment of the engine it is also beneficial to cool thecrankcase by cooling the oil in the crankcase. An oil cooler 10941 showndiagrammatically in FIG. 109B is designed to pick up a substantialamount of the heat generated in the crankcase, and with a co-axial (or atube-in-tube) heat exchanger 11043 shown specifically in FIGS. 110A-B,oil from the crankcase passes through an outer oil channel 11045 over aseries of fins 10947 positioned along the outer surface of a coolingtube 11049 containing flowing cooling water from a cool water source11046. The fins 11047 can be radial fins or axially aligned finsrelative to the cooling tube 11049 depending upon the necessity for adesired oil flow along the outer surface of the coolant tube. Aftertaking up heat from the oil, the cooled oil returns to the main line11037 and the heated water can be dumped to a heat sink 11051.

A manual pull-cord start and recoil pulley assembly may be added to theengine to assist in engine start-up in other embodiments. FIGS. 111A andB show a pull-cord start system 11193 including a recoil pulley 11195coupled to a crankshaft 11197 supported preferably in a Stirling enginecrankcase. A pull-cord 11115 is wound about the recoil pulley 11195which is pulled by a user to rotate the recoil pulley and thereby thecrankshaft 11197 at a speed sufficient to start the engine. The pulleyis preferably surrounded by a stationary housing (not shown) for safetyreasons and is coupled to the crankshaft via a pulley shaft 11109 and aone-way clutch 11119 that drives the crankshaft as the pull-cord 11115is pulled and permits the crankshaft to freely rotate relative to thepulley 11195 when the engine is running.

When starting the engine, the operator manually grasps the handle Hattached to the pull-cord 11115 and pulls the pull-cord 11115 outwardfrom the recoil pulley 11195. This turns or rotates the pulley in eithera counter-clockwise or clockwise direction as shown in FIG. 111A againstthe bias of the torsion spring (not shown) generally engaged between thepulley and the housing. The operator must pull the pull-cord 11115 withsufficient strength to overcome the bias of the torsion spring thatwould normally cause the pull-cord 11115 to rewind back into the housingand around the circumferential groove in the pulley. As the pull-cord11115 is pulled outward the pulley 11195 engages through the pulleyshaft and one-way clutch 11119 with the crankshaft 11197 of the enginecausing the pistons (not shown) to reciprocate with sufficient speed tostart the engine. When the pull-cord 11115 is released by the user, therecoil spring (not shown) causes the pulley 11195 to rotatecounter-clockwise through a series of complete revolutions. Where theend of the pull-cord 11115 is connected to the pulley, when the userreleases the handle H after pulling the pull-cord, the pull-cord 11115travels or rewinds on the pulley 11195 and recoils back inside thehousing placing the recoil start up assembly back into a recoiled state.

Alternatively as shown in FIGS. 111C and 111D an electric starter motormounted externally to the engine housing may be used to start theStirling engine. The electric motor 11196 may be a permanent magnet or aseries-parallel wound direct current electric motor with a solenoidswitch 11192. When current is applied to the solenoid 11192, a drivepinion or spur gear 11194 is pushed out along a driveshaft 11198 to meshwith a ring gear 11191 that is coupled to the crankshaft 11197. Thedriveshaft 11198 may be a helical Bendix drive that drives the pinion11194 along the helix to engage with the ring gear 11191. When theengine starts, backdrive from the ring gear 11191 causes the drivepinion 11194 to exceed the rotative speed of the starter motor 11196 andforce the drive pinion 11194 back down the helical shaft and thus out ofthe mesh of the ring gear 11191. The drive pinion 11194 mayalternatively be clutched to its drive shaft through an overrunningsprag clutch which permits the pinion 11194 to transmit drive in onlyone direction. If the pinion 11194 remains engaged after the engine hasstarted, the pinion 11194 will spin independently of its driveshaft 1119preventing backdrive that causes the engine to drive the starter andpossibly damage the starter motor 11196.

B-Burner

FIGS. 112-119 disclose a further embodiment of a burner 11201 for use inconjunction with a multiple heater head and piston engine describedpreviously in FIGS. 90-91B. The present burner 11201 is specificallydirected to the independent heating of multiple heater heads, in thiscase four (4) heater heads, each heated by an individual burner andflame and having a single air inlet 11223, single outer wall 11212, andtwo exhaust openings 11225.

Turning to FIG. 113 the four burner design 11301 of the presentembodiment includes a single blower B providing air for the fuel/airmixture in the ignition process of all the burner head assemblies 11305as shown in FIG. 113. The heater heads 11303 as also discussed above,may be any of the various embodiments described in the precedingsections, including, but not limited to, tube heater heads, asdesignated by numeral 9116 in FIGS. 91B-91D, or pin or fin heater heads,as designated by numeral 9118 in FIG. 91A (and also shown as 5100 inFIGS. 53D through 53F). By way of example, the present embodiment iscontemplated utilizing heater tubes 11309 through which flow the workinggas, for example helium, which must be heated by the burner headassemblies 11305 during the appropriate portion of the Stirling cycle.

By way of more detailed description and referring back to FIG. 112 aswell, the burner 11301 includes multiple burner head assemblies 11305,one for each of the heater heads 11303, in the case of the presentembodiment there are four (4) heater heads and hence four (4) burnerhead assemblies 11305. The cross-section of FIG. B shows three (3) ofthe burner head assemblies 11305. Generally, the burner 11301 is definedby a burner housing as shown in FIG. 112 having a substantiallycylindrical outer wall 11312, although other geometrical configurationscould be imagined. The blower B pumps air into the burner 11301 throughair intake 11223 for purposes of ignition and combustion, and exhaustgases are ejected from the burner via the two exhaust outlets 11225adjacent the base of the burner.

Turning to FIGS. 114 and 115 a top surface 11413 of the burner housingincludes a number of ports 11415 for receiving fuel inputs 11416,igniters 11427, flame and possibly temperature sensors or flame viewingelements. The ports 11415 also facilitate access to a particular burnerhead 11405, as discussed in detail below, without having to remove theentire burner 11401 from the vessel stack-up for maintenance. As seen inFIG. 115, associated with each burner port 11515 on the top surface11513 of the burner is a secondary port 11517 which can serve a numberof purposes for instance a flame viewing element such as a viewingwindow for viewing the flame of the burner head, or alternatively aspark plug 11520 for igniting the fuel/air mixture and/or a sensor forsensing UV light used in flame detection.

The base of the burner 11601 best seen in FIGS. 116 and 117, is providedwith heater head openings 11619 to accommodate the entrance of theheater heads and respective heater tubes since the burner as a whole isset over and stacked up on a cooling plate 11604 of the vessel so thatthe heater heads 11603 are within and substantially sealed inside orencompassed by a lower region of the burner 11601. The base of theburner 11701 is secured to the mounting plate 11704 in the vesselstack-up by a circumferential band clamp 11710, such as a Marmon clamp,which is provided for securing and critically circumferentiallycentering the burner relative to the cooling plate and lower stack-up ofthe pressure vessel. The centering of each burner head assemblies 11705relative to each of the associated heater heads 11703 is criticalbecause if the flame from the burner head assembly is nearer one side ofthe heater heads 11703 and heater tubes 11709 than another, there willbe not only inefficient heating of the working gas/fluid in the heatertubes 11709, where one set of heater tubes is heated to a highertemperature than other tubes.

The clamp 11710 extends circumferentially and radially around the entirebase of the burner 11701 and provides both radial and axial compressiveforces between the burner base plate and the mounting plate to ensurethat there is both a critical axial sealing pressure to contain the hotexhaust gases in the burner and a radial circumferential alignment ofthe burner heads with the heater heads. The base of the burner housing11711 may be provided in this regard with a circular sealing edge 11721as shown in FIG. 117 which is angled relative to the vertical axialarrangement of the vessel stack-up to create the axial compressive forceand for mateably engaging with an oppositely angled circular sealingedge 11722 of the mounting plate. The circumferential clamp 11710 andthe mating angled circular sealing edges 11721,11722 of both the burnerand the cooling plate ensure the critical circumferential, i.e. radialalignment of the burner housing 11711 and burner head assemblies 11705with the mounting plate and heater heads 11703 in the vessel stack-up sothat the burner head assemblies 11705 are appropriately aligned with theheater heads 11703 and there is sufficient axial force between theburner 11701 and cooling plate 11704 to contain the hot exhaust gasgenerated in the burner 11701. The circular sealing edge 11721 mayinclude a graphite seal (not shown) between the burner and cooling plateto ensure that the hot gases which are at around 1000 C, where the flametemperature is around 1200 C, do not leak out between the burner 11701and the mounting plate 11704.

The single blower B as shown in FIG. 117 provides air into the burnerhousing 11711 adjacent the top surface 11713 for the fuel/air mixture tothe burner head assemblies. The air intake 11723 is provided atessentially a normal angle to the circular burner housing 11711 andprovides air inside the burner 11701 for combustion as described indetail below. This arrangement of the air intake 11723 at a normal angleto the cylindrical burner housing 11711, better seen in FIG. 116,facilitates air entering the burner 11701 with a designed pressure dropwhich is important for incoming air to the burner to maintain a desiredair velocity for maximizing heat transfer efficiencies as the air passesthrough the air intake manifold and into a preheater 11726 where theincoming (cold) air is warmed by the exiting (hot) exhaust gasses. Asingle blower B is placed in communication with the single air intake11723 to provide air to all the burner head assemblies 11705 in theburner 11701. A pair of exhaust outlets 11725 are also connected normalto the substantially cylindrical burner 11701 and spaced approximately180 degrees apart around the base of the burner 11701. Prior to exitingthrough the exhaust outlets 11725 the exiting exhaust from the burner11701 preheats the incoming air in the preheater 11726 described indetail below, then exits the burner 11701 from one of the two exhaustoutlets 11725.

Observing FIG. 117, each burner head assembly 11705 has a fuel injector11724, an igniter 11727 of one kind or another, for instance a sparkplugor glow-plug, a flame detection device 11729 which may also be providedin the secondary port 11717 as shown. Fuel, either liquid fuel orgaseous fuel is fed to the fuel injector 11724 via a fuel line 11731from a fuel source F and is dispersed as a fine mist or vapor by thenozzle 11734 of the fuel injector 11724 into a prechamber 11728 of theburner head assembly 11705. In the prechamber 11728 the dispersed fuelis combined with a desired volumetric flow of air from the preheater11726, preferably preheated to a desired ignition temperature by theexhaust as discussed in detail below, to form a desirable fuel/airmixture for ignition. The fuel/air mixture then is ignited by theigniter 11727 and combusts at least partly inside the prechamber 11728,but more complete combustion may occur after the fuel/air mixture exits,or is pushed, from the prechamber 11728 through the prechamber nozzle11730 of the prechamber 11728 to form a flame which extends from theprechamber 11728 and is directed into a center combustion chamber insidethe heater tube arrangement of each respective heater head 11703.Exhaust from the combustion in the burner 11701 exits the burner via thepreheater 11726 and exhaust manifold 11714 described in detail below.

In the present embodiment of the burner, the single blower B, shown herediagrammatically, may be incorporated to maintain a consistent averageair ratio supplied to the burner 11701 and hence to each of theindividual burner head assemblies 11705. The blower B pumps air at adesired velocity depending on instructions from a controller C forpurposes of ignition, then once ignition has occurred, the desired airflow rate may be regulated by the controller C dependent on datareceived from sensors including but not limited to an oxygen sensor. Amore complete description of the burner control algorithm is providedbelow. The single blower B is also controlled dependent on the data fromindividual burner head assemblies, for example in the case of at leastone burner head assembly being extinguished or not igniting thecontroller may decrease the blower rate to facilitate ignition in theextinguished burner head assembly. The fuel input may be correspondinglycontrolled in the remaining burner head assemblies 11705 to accommodatesuch an air velocity decrease. In any event, the blower B is intended toprovide a consistent flow rate to each of the multiple burner headassemblies 11705 in the burner after passing through the preheater11726. An important aspect of the present embodiment is the consistentflow and velocity of cold air developed by the blower B and theefficient heating of the incoming air through the extraction of wasteheat from in the preheater 11726, to raise the cold air temperaturethereby improving the efficiency of combustion processes and the burnerunit.

The blower B connects through the air intake 11723 in the outer wall11712 of the burner housing into a cold air channel 11735 of thepreheater 11726. The cold air channel 11735 extends circumferentiallyaround the burner inside the outer wall 11712 of the burner 11701 anddirects the cold air developed by the blower B down around an insulatedintermediate baffle 11739 and up into the preheater. The intermediatebaffle 11739 is insulated to protect the outer wall 11712 of the burner11701, and anyone or thing that comes in contact with the outer wall,from the intense high temperatures inside the burner 11701. Also, theinsulated baffle 11739 ensures that heat captured by the incoming air inthe preheater 11726 is not lost directly to the outer wall 11712 of thehousing 11711.

The preheater 11726 essentially begins where the cold air from theblower B drops down through the cold air channel 11735 and enters into apreheater channel 11741 in which the cold air is preheated in order toraise its mean temperature which increases the efficiency of the burnercombustion. The preheater channel 11741 is defined by the intermediatebaffle 11739 on one side, and on the inner side, an exhaust manifoldwall 11743. The exhaust manifold wall 11743 directly separates theincoming cold air from the exhaust air exiting the burner and providesfor the heat transfer from the exiting exhaust to the incoming cold airin the preheater channel 11741. The heat transfer efficiency through themanifold wall 11743 in the preheater 11726 is critical because thehotter the incoming cold air can be raised by the preheater 11726, theless fuel is necessary to get the gas up to desired ignition andcombustion temperatures. The preheater channel 11741 also extendscircumferentially around the entire burner 11701 which provides for amaximum surface are which in some embodiments may produce better heatexchange with the exhaust flowing out of the burner through an exhaustchannel 11744. Inside the preheater channel 11741 are a series ofradially extending fins 11745 which are directly connected to theexhaust manifold wall 11743 and assist in efficient heat transfer fromthe exiting exhaust air through the manifold wall 11743 into the air inthe preheater channel 11741. Exhaust side fins 11746 may also beconnected to the exhaust manifold wall 11743 extending into the exhaustchannel 11744

The cold incoming air is preheated to a desired temperature, forexample, but not limited to 600-750 C, in the preheater 11726 whichfacilitates ignition and combustion as the air is directed to the burnerhead assemblies. The amount of preheating which may be accomplished isprimarily based on the efficiency of heat transfer from the exitingexhaust so that as the exhaust temperature is raised during operation ofthe engine, the incoming cold air can be accordingly preheated to ahigher temperature. The preheated air exits the preheater channel 11741and is directed radially into a hot air chamber 11747 which communicateswith each of the multiple burner head assemblies 11705. It is to beappreciated that the preheated air enters the hot air chamber 11747through a substantially 360 degree circumferential opening around theexit of the preheater channel 11741 so that a consistent flow rate ofpreheated air is delivered to each of the burner head assemblies 11705.While additional channels or passageways (not shown) may be provided inthe hot air chamber 11747 to direct the preheated air in the hot airchamber to a specific burner head, the 360 degree output from thepreheater channel of the present embodiment is important since there isonly one blower B developing the air flow into the engine. In previousengines a multitude of blowers delivered a desired air flow to each ofthe burner head assemblies, for instance where there were four (4)burner head assemblies, there were four (4) blowers, one directed toeach burner head. However, having a blower associated with each burnerhead 11705 on a multiple burner head engine is expensive and adds asignificant amount of weight to the engine. In any event, a singlepreheater is much less expensive and less complicated from a controlstandpoint than separate preheaters of each heater head.

The preheated air is directed in the hot air chamber 11747 to theindividual burner head assemblies 11705 and specifically to anintersection with a nozzle 11734 of each fuel injector 11724 in eachburner head 11705 and an igniter 11727. The fuel injectors 11724 may useeither liquid fuel or gaseous fuel but in either case the fuel isejected from the injector into the prechamber 11728 where the fuel mixeswith the preheated air to attain a desired fuel/air ratio or mixture foreither ignition of the burner head 11705, or, combustion where theburner head 11705 is currently supporting a flame. The fuel injectors00024 inject the fuel into the prechamber 11728 directly below the fuelinjector 11724 and the preheated air is combined in the prechamber 11728with the liquid or gaseous fuel. The fuel may be delivered as a mist orvapor, combined with the preheated air and ignited in the prechamber11728 by the igniter 11727. While ignition of the fuel/air mixture mayoccur to some extent in the prechamber 11728, the flame derived from theignition and combustion of fuel/air in the prechamber 11728 needs to bepushed out of the prechamber 11728 to be more efficient and provide therequisite thermal output. It is preferable that the constant combustionflame which heats the heater heads 11703 and heater tubes 11709 bepushed out of the prechamber 11728 and actually extend beyond the endcone 11730 of the prechamber 11728 and into the combustion chamber11750. This is accomplished by providing appropriate adjustment to thefuel/air mixture by the controller and by the prechamber and nozzlegeometry to properly control the shape of the flame.

The shape of the flame emanating from the prechamber 11728 is importantto the efficiency of the engine. The exit cone 11730 and other aspectsof the prechamber 11728 itself can be shaped, sized and arranged tocontrol the shape of the flame emanating from the prechamber. It isimportant to attain a desired shape of the heating flame since an evenheating of the heater tubes 11709 and heater heads 11703 is highlydesirable. A preferred shape and dynamic motion to the flame is an axialrotation of the flame in the combustion chamber 11750 which develops anevenly distributed flame and heating of the heater tubes 11709 in thecombustion chamber. To assist in obtaining an appropriate axial rotationof the flame a fuel/air dispersion element(s) can be provided in theprechamber. Such a dispersion element could be an insert like afletching 11849 shown in FIGS. 118A-D, positioned inside the prechamber11828, defining essentially fins, channels or baffles around an insidesurface of the prechamber. Although other and different shapes arepossible, the helical fletchings 11849 can be formed or attached forinstance by welding on the inside wall of the prechamber 11828 as shownin FIG. 119. These fuel/air dispersion element(s) inside the prechamber11828 help control the rotation of igniting gases and fuel in theprechamber 11828 in such a way as to attain a desired flame shape as itemanates from the prechamber 11828 and into proximity with the heatertubes 11809. The fletchings 11849 can be shaped in different ways andeven two oppositely disposed fletchings can be affixed inside theprechamber to initiate and direct the axial rotation of the fuel/airmass in the prechamber 11828. The fuel/air flame shape is controlled bysuch fletchings 11849 for example to define the desired symmetricalshape and axial rotation of the output flame on the heater head. Themore even and symmetrical the flame out of the prechamber 11828, themore efficient and even heat dispersion occurs and is accorded to theheater tubes on the heater heads.

Another aspect of the present embodiment shown in FIG. 119 is theprechamber support which extends around the outermost wall of theprechamber 11928 adjacent the exit cone 11930 and includes a heater headrestrictor 11937 restricting the flame exhaust gases from passing overthe heater head 11903. The restriction increases the pressure drop atthe top of the heater head 11903 and improves heat transfer from the hotgas to the helium inside the heater tubes 11909 by encouraging the hotgas to flow through the heater tubes.

A substantial amount of the heat not used to heat the working fluidremains in the exhaust gases and thus the efficiency of the entireengine can be increased by using the excess exhaust gas heat to preheatthe incoming air. After heating the heater head 11903 and heater tubes11909 the hot combustion gases are forced out an exhaust inlet 11953 bynewer combusted gases into the exhaust channel 11944 defined partly byan inner wall 11942 of the exhaust manifold. The exhaust passes downalong the exhaust channel 11944 exchanging a substantial amount of heatthrough the preheater wall 11943 to the incoming cold air entering thepreheater 11926 via the preheater channel 11941. The exhaust gasesshould flow as quickly as possible through the preheater 11926 as theheat transfer from the exhaust gases is dependent upon the velocity ofthe exhaust. Another aspect of the present embodiment limits thepressure drop of the exhaust gases by allowing the exhaust gases flowout from two exhaust outlets 11925, as opposed to one exhaust outlet,arranged around the bottom of the exhaust manifold 11914. The shorterflow path provided by the two exhaust outlets 11925 for the exhaustleaving the exhaust manifold 11914 lowers the pressure drop for theexhaust, the blower does not have to work as hard, and thus the blowerload is reduced on the engine.

FIG. 119 also discloses the use of spark plugs 11960 in the secondaryport 11917 (rather than a UV viewing window) through the burner housing.In certain cases a flame sensor may also be inserted through thesecondary port 11917 which extends into the burner adjacent each of theprechamber nozzle so that flame detection can occur. In any event, thesecondary port 11917 provides for access to the burner head 11905 sothat a sensor or window, or ignition components for instance glo-plugsor spark plugs as shown here can be inserted down into the burner head11905 to ignite different gases and fuels. Gaseous fuel use maynecessitate the spark plug 11960 to ignite the fuel air mixture adjacentthe nozzle of the prechamber whereas liquid fuel uses glow plugs and aregenerally located closer to the fuel injector itself. In the embodimentshown here a high voltage conductive element 11961 is encased within andinsulative layer 11963 and a ground layer 11965 and inserted through thesecondary port so that the exposed conductive element 11961 is exposedin the combustion chamber to ignite the fuel/air mixture exiting theprechamber 11928.

The ability to see and/or detect each flame is important so that each ofthe four individual burner head assemblies 11905 and respective flamecan be appropriately adjusted by the controller. It is to be appreciatedthat such flame detection and viewing may be accomplished by manyembodiments, including but not limited to an actual viewing window forexample having appropriate lenses in the tube which allow a humanoperator to look through the tube and visually identify a flame withinthe range of visible wavelengths in the combustion chamber.Alternatively, the viewing window may include a camera or other imagedata receiving and recording device such as a UV light sensor anddisplay for visually displaying a received representation of the flamein the combustion chamber. Other types of heat sensors including but notlimited to thermocouples, infrared thermometers, and thermisters, may beused to identify and quantify the flame and flame characteristics in thecombustion chamber.

With only a single blower providing air to four burner head assemblies11905, generally a variable in addition to air, such as fuel, must bealtered to obtain a desired flame quality. Keeping one blower providingair to all four burner heads is especially helpful for cost and forblower power consumption.

With liquid, diesel or other gaseous fuel, the UV viewing window will becompromised because the fuel vapor tends to absorb the UV radiation fromthe flame. Without the UV window as in the previous embodiment it maystill be important to detect the flame and the temperatures in thecombustion chamber. The electrode of the sparkplug may be utilized as asensor in some cases to detect the flame. Such data can be forwarded tothe controller to determine the flame and combustion disposition in thecombustion chamber 11950. Another method of flame detection obtainstemperatures with temperature sensors inside the heater head, forexample a thermocouple attached on the walls of the heater tubes canprovide data to the controller to alter the operational conditions ofthe engine. This temperature data is used to judge the temperatureand/or flame quality based on temperature/flame data and helps thecontroller decide what operational mode, as discussed in further detailbelow, to set for each burner head 11905 and for the engine as a whole.

Burner Control

The burner may be operated in several modes as shown in FIG. 120, andreferring in part to FIG. 119, according to predetermined electronicsand software programs embodied in an electronic controller. Theoperation modes evaluated by the controller include at least a start-upmode 12002, normal operation mode 12004, shut-down mode 12006 and a stopmode 12008. The start-up mode includes the initial ignition of a richerfuel mixture in the prechamber 11928 to ease ignition as colder mixtureshave a narrower ignition range of fuel/air ratio that are ignitablecompared to the range of fuel/air ratios that maintain combustion. Witha desired fuel/air ignition mixture present in the prechamber 11928, theigniter 11927 is actuated and the ignition mixture is ignited. Athermocouple (not shown) in the prechamber 11928 detects what isreferred to as a diffusion flame in the prechamber 11928 and once theincoming air is hot enough from the preheater 11926, the flame is pushedout of the prechamber 11928 by either increasing the air flow from theblower B, or increasing fuel so the flame travels out of the prechamber11928 and forms in the combustion area adjacent the heater head 11903.

Generally in the start-up mode 12002 as shown in FIG. 120 a user sets adesired blower speed 12003 and fuel/air ratio 12005 for a certain timeperiod 12007, for example 30 seconds. After the predetermined timeperiod the blower shuts off and resets 12011 the start-up phase whichmay include blowing out 12013 any remaining fuel in the engine andexhaust system so that there are no backfires or other damaging eventsfrom residual fuel. The start-up phase may also include for instance anumber of ignition attempts 12009 before resetting and providing theuser with an error A sensor (not shown) within the prechamber 11928 or avisual sensor using the secondary port 11917 detects if a flame 12010 ispresent within the prechamber 11928 or the combustion chamber 11950. Ifa flame is not detected the system is reset 12011 or if a flame isdetected the temperature readings are taken 12015 from the heater headand oxygen levels are measured 12017 from the exhaust gases. Thefuel/air ratio is then adjusted 12019 based on these readings.

Once the flame is supportable out of the prechamber 11928 and is heatingthe heater head 11903, the control system and operation mode 12004include a number of failsafe triggers 12023 based on sensor data andcontroller evaluation algorithms which evaluate the system and determineif the system should be turned to the shut-down or stop mode. Theoperation mode 12004 monitors levels of heat, power and oxygen forexample and perform shut-down or stopping of the engine, or othermodifications to the system and engine if a temperature reading is toohigh, or exhaust oxygen level is too high or if engine speed exceeds adesired value, or the differential pressure within the air lock is toolow. These are just exemplary triggers for starting shut-down or stopprocedures, other triggers could be used as well or in combination withthese examples.

During normal engine operation, the blower is operated at leastpartially by a control loop which measures the excess oxygen 12017 inthe exhaust to determine blower speed. The failsafe triggers 12023 shownin the flowchart and operation analysis table 12021 in FIG. 120 include:Engine speed exceeds predetermined range; Oxygen levels in exhaustexceed a predetermined range; Generator temperature exceeds apredetermined range; Burner temperature exceeds a predetermined range;Cooler temperature exceeds a predetermined range; Flame/Ignitionfailure; repeatable Failure of flame ignition. It is to be appreciatedthat the described control method is not limited to the disclosedtriggers 12023 and that other triggers, factors and variables may alsobe analyzed by the controller under the start-up and operation modes12002 and 12004.

A failure of the engine in one of these failsafe triggers 12023 directsthe controller C to adjust the fuel/air ratio 12019 and continueacquisition of sensor readings. A preset number of a repeated failure12025 of the engine to run within a predetermined range for any of thesetriggers leads to a shutdown sequence with an immediate fuel turn off12029. The engine however can continue to run in the shut-down mode12006 in many cases. On the other hand, certain events may causecomplete engine stoppage (i.e. shut-off as opposed to shut down) so thatdamage to the engine is minimized. A status check 12037 on systemcomponents is repeatedly run. These shut-off triggers 12034 are forexample, low oil pressure, low airlock pressure differential, and lowengine power levels will ensure complete engine stoppage to preventdamage.

During a shut-down mode 12006, the fuel and burner is turned off but theengine keeps running until the heater head 11903 is cooled to a desiredtemperature. A system shut-down may also be caused by excessive heatmeasurements in a number of components such as the Generator, theburner, or a cooler, or a system shut down may occur if there is afailure to ignite. A shut down due to system failure may trigger a safemode where fuel is pumped out of the system. Any fault or system failureor trigger, will kill the fuel delivery immediately 00036, but theengine will continue to run to cool down the system. The engine runsuntil it reaches a predetermined power level 12035 in the shut down mode12006, or in the event of the more dangerous fail safe triggers theengine is stopped 12008, i.e. the RPMs are set to 0. The shut-down modehelps engine efficiency since the engine, burner and heater heads remainhot for a while, even while there is no fuel supplied, the engine willcontinue to run producing power until the predetermined low power levelis reached. This recovers some of the energy put in at start-up modewhich improves efficiency.

C Burner

A still further embodiment of the present invention is a multiple pistonengine with a single central burner as disclosed in FIGS. 121 and 122.From a process control standpoint the central burner 12101 facilitatescontrol of the ignition, flame temperature and shape, pressure drop andfuel air ratio for a single burner, single blower, single fuel injectorand flame. This can greatly simplify the above described burnercontrols, hardware and accessories. The single burner design depends inlarge part on an engine and system design which provides highlyefficient and controllable transfer of heat from the central combustionchamber 12250 to multiple heating chambers 12118 where the heater heads12203 are located so that the working fluid in the heater head tubes12209 absorbs the appropriate heat.

The single burner 12101 embodiment of the presently described engineincludes a single burner head, where the heat from one combustionprocess, i.e. one flame, is dispersed to all the heater heads within theheating chambers, for example to four (4) heater heads and best seen inFIG. 122. Structurally somewhat similar to the previously describedembodiment, burner 12101 includes a burner housing 12111 having asubstantially cylindrical outer wall 12112, although as noted aboveother geometrical configurations could be imagined, and a top surface12113 defining a number of ports for receiving various other elementssuch as, air, fuel lines, flame and temperature sensors or flame viewingelements. A central port 12115 permits access to the central burner headand remaining ports 12116 facilitate access to a heating chamber.

In addition to the central port 12115, the top surface of the burner12105 may have a secondary port 12117 which can serve a number ofpurposes for instance a viewing window for viewing the flame of theburner head and a spark plug for igniting the fuel/air mixture and/or asensor for sensing heat and flame temperature. The secondary port 12117may also include a tube for the addition of air/oxygen and/or a sensorsuch as a temperature sensor for sensing flame temperature. The heatingchamber ports 12116 may have an air inlet passage 12132 to providecooling air to the heater chambers 12118. As will be discussed below,cooling or additional air may be added to the heater chambers 12118through the air inlet passages 12132 to help control the temperature ofthe heating chamber 12118 and the heater heads 12203 during engineoperation.

Turning to FIG. 122, the prechamber 12228 is arranged above a combustionchamber 12250 that is positioned centrally between all the multipleheater heads 12203, i.e. the heater heads and tubes are outboard of thecombustion chamber 12250. The heater heads 12203 and heater tubes 12209are in the heating chambers 12218 external to the combustion chamber12250 as will be described in detail below. The base of the burner 12201is provided with openings 12219 to accommodate the entrance of theheater head 12203 and heater tubes 12209 since the burner as a whole isset over and stacked up on the cooling plate 12204 of the vessel so thatthe heater heads 12203 are within and substantially encompassed by alower region of the burner 12201.

The base of the burner is secured to the cooling plate 12204 in thevessel stack-up by a circumferential band clamp 12210, such as a Marmonclamp, as described above for securing and critically circumferentiallycentering the burner 12201 relative to the cooling plate 12204 and lowerstack-up of the pressure vessel. In centering the burner 12201, thecombustion chamber 12250 of the single burner head 12205 is centeredrelative to each of the associated heater chambers 12218 and heaterheads 12203. The centering of the combustion chamber 12250 is criticalbecause the heat transfer from the hot exhaust gases from the combustionchamber 12250 to the heating chambers 12218 must not be nearer to oneheater head 12203 than another within the heating chamber 12218, becauseheating of the working gas/fluid in the heater tubes 12209 must be evenfor better engine efficiency.

As described above, the clamp 12210 extends circumferentially andradially around the entire base of the burner 12201 and provides bothradial and axial compressive forces between the burner housing 12211 andthe cooling plate 12204 to ensure that there is both a critical axialsealing pressure to contain the hot ignition and exhaust gases in theburner 12201 and a radial circumferential alignment of the combustionchamber 12250 of the burner head 12205. The base of the burner housingas noted above may be provided in this regard with a circular sealingedge 12221 which is angled relative to the vertical axial arrangement ofthe vessel stack-up for mateably engaging with an oppositely angledcircular sealing edge 12222 of the cooling plate 12204. Thecircumferential clamp 12210 and the mating angled circular sealing edge12221 of both the burner 12201 and the cooling plate 12204 ensure thecritical alignment of the burner housing 12211 with the cooling plate12204 and heater heads 12203 in the vessel stack-up so that thecombustion chamber 12250 of the burner head 12205 is appropriatelyaligned and there is sufficient axial force between the burner 12201 andcooling plate 12204 to contain the hot gas and exhaust generated in theburner 12201. The circular sealing 12221 edge may include a graphiteseal between the burner 12201 and cooling plate 12204 to ensure that thehot gases which can be around 1000 C where the flame temperature isaround 1200 C do not compromise the seal between the burner 12201 andthe cooling plate 12204.

Similar to the multiple head burner described above, a single air intake12223 is provided through the burner housing 12211 adjacent the topsurface 12113 and the ports 12115, 12116 to deliver air for the fuel/airmixture to the burner head 12205. The air intake 12223 is connected to ablower B and defines an air intake passage 12235 that is provided atwhat is essentially a normal angle to the circular burner housing 12211.As noted above, this arrangement provides for the designed or desiredpressure drop for incoming air to the burner 12205 as it passes throughthe air intake passage 12235 and into the preheater 12226 where theincoming air is preheated by the exiting exhaust air. The single blowerB is placed in communication with the air intake 12223 to provide air tothe central burner head 12205 in the burner 12201 and a pair of exhaustoutlets 12225 are also connected normal and spaced approximately 180degrees apart around the base of the burner 12201. The exiting exhaustfrom the burner 12205 preheats the incoming air from the blower B in thepreheater 12226 as described above then exits the burner 12201 aftercombustion occurs from one of the two exhaust outlets 12225.

The burner head 12205 as shown in FIG. 122 has a single fuel injector12224, an igniter 12227 of one kind or another for instance a sparkplugor glow-plug, and a flame detection device 12229. Fuel, either liquidfuel or gaseous fuel is fed to the fuel injector 12224 via a fuel line12231 from a fuel source F and is dispersed as a fine mist or vapor bythe nozzle 12234 of the fuel injector into a prechamber 12228. In theprechamber 12228 the dispersed fuel is combined with a desiredvolumetric flow of air from the blower B through the preheater 12226,preferably preheated to a desired temperature by the exhaust asdiscussed in detail above, to form a desirable fuel/air mixture forignition. The fuel/air mixture is then ignited by the igniter 12227 andcombusts at least partly inside the prechamber 12228 but more completelyafter exiting the prechamber 12228 through the exit cone 12230 of theprechamber 12228 to form a flame which extends into the centercombustion chamber 12250. The top of the combustion chamber 12250 hasoutlet channels which direct the hot gas from the combustion into theheating chamber 12218. Exhaust from the heater chamber 12218 in theburner exits the burner via the exhaust inlet 12253 and exhaust manifold12214 described in detail above. While a description of the presentembodiment includes the central combustion chamber 12250 in the burneras provided above, it is to be appreciated that other embodimentscontemplate additional burners for example a two burner head design witheach burner heating a pair of heater heads where such an arrangement maynot have centrally aligned burner head assemblies and combustionchambers relative to the burner itself.

In the present embodiment of the burner, the single blower B may beincorporated to maintain a consistent average fuel/air ratio to theburner head 12205. As noted above, the blower B pumps air at a desiredrate depending on various data from the controller for purposes ofignition, then once ignition has occurred, the desired air flow rate maybe dependent on the measured residual oxygen in the exhaust, however insome embodiments it may be dependent on at least the desired temperatureof the burner and a desired power output of the machine. The singleblower B is also controlled dependent on the data from burner head 12205and the heating chamber 12218, for example in the case of the burnerhead being extinguished or not igniting the controller may increase ordecrease the blower power to facilitate re-ignition.

The blower B is intended to provide a consistent air flow rate throughthe preheater 12226 to the burner head 12205. An important aspect of thepreheater 12226 of the present embodiment is the channeling of theconsistent flow and velocity of cold air developed initially by theblower B in the preheater channel 12226 across the exhaust manifold wall12243. Because the blower inputs a consistent flow of cold air into theburner and through the preheater 12226, i.e. the heat exchanger, theefficiency of the burner is increased because more waste heat isextracted from the exhaust to heat the incoming cold air.

As described above, the blower connects through the air intake 12223 inthe outer wall of the burner housing 12211 into a cold air channel12235. The cold air channel 12235 essentially defines the entrance intothe preheater 12226 and extends circumferentially around the burnerinside the outer wall of the housing 12211 and directs the cold airdeveloped by the blower B down around an insulated intermediate baffle12239 and up into the preheater 12226. The intermediate baffle 12239 isinsulated to protect the outer wall of the burner, and anyone or thingthat comes in contact with the outer wall, from the intense hightemperatures inside the burner. More importantly, the insulated baffle12239 ensures that heat captured by the incoming air in the preheater12226 is not lost directly to the outer wall of the housing. This alsoensures that the incoming air is cold which improves the preheaterefficiency

The preheater 12226 of the burner 12201 has been described in detailabove with respect to the previous burner embodiment and essentiallyincorporates the same or similar channels and baffles to direct cold airalong the exhaust manifold wall 12243 to facilitate direct heating ofthe incoming air as previously described. The cold air is preheated to adesired temperature which facilitates ignition and combustion as itrises through the preheater channel 12241. The amount of preheatingwhich can be accomplished is based on the heat transfer from the exitingexhaust so that as the exhaust temperature is raised during operation ofthe engine, the incoming cold air can be accordingly preheated to ahigher temperature. The preheated air exits the preheater channel 12241and is directed radially into a hot air chamber 12247 which communicateswith the single burner head 12205. It is to be appreciated that thepreheated air enters the hot air chamber through a substantially 360degree circumferential opening around the exit of the preheater channel12241 so that a consistent, uniform flow rate of preheated air isdelivered to the burner head. Channels or fins (not shown) may beprovided in the hot air chamber to direct the preheated air in the hotair chamber to the burner head.

The preheated air is directed in the hot air chamber 12247 to the burnerhead 12205 and specifically to an intersection with a nozzle 12234 ofthe fuel injector 12224 in the single burner head and an igniter 12227.As previously described, the fuel injector 12224 may deliver eitherliquid fuel or gaseous fuel but in either case the fuel is ejected fromthe injector into the prechamber 12228 where the fuel mixes with thepreheated air to attain a desired fuel/air ratio or mixture for eitherignition of the burner head or combustion where the burner is currentlyproducing a flame. The fuel injector 12224 injects the fuel into theprechamber 12228 and the preheated air is combined in the prechamber12228 with the liquid or gaseous fuel. Also as noted, the fuel may bedelivered as a mist or vapor, combined with the preheated air andignited in the prechamber 12228 by the igniter 12227. While ignition ofthe fuel/air mixture may occur to some extent in the prechamber 12228,the flame derived from the ignition and combustion of fuel/air in theprechamber 12228, it is preferable that the constant combustion flamewhich heats the heater chambers 12218 and heater heads 12203 and tubes12209 be pushed out of the prechamber 12228 and actually extend beyondthe end cone 12230 of the prechamber 12228 and into the combustionchamber 12250.

In the present embodiment with only a single burner 12205 and the centercombustion chamber 12250 where the flame is not directly adjacent thetube heaters 12209 in the heater heads 12203, the shape and structure ofthe flame is not as critical where the flame is not in contact with theheater heads. Just like the previous embodiments though, the end cone12230 and even other aspects of the prechamber 12228 itself can beshaped and sized and arranged to control the shape of the flameemanating from the prechamber 12228. It is still important to attain adesired shape to the heating flame since an even dispersal of heat fromthe combustion chamber 12250 to the heater chambers 12218 containing theheater tubes 12209 and heater heads 12203 is highly desirable. In otherembodiments the inner combustion chamber may be used to recirculate andre-burn some of the exhaust gas to reduce emissions.

As in the previous embodiment it may be preferable to impart apredefined shape and dynamic motion to the flame for an axial rotationof the flame in the combustion chamber 12250 which develops an evenlydistributed flame and heating of the heater tubes 12209 in thecombustion chamber 12250. To assist in obtaining an appropriate axialrotation of the flame a fuel/air dispersion element(s) can be providedin the prechamber 12228. As previously discussed such a dispersionelement could be a centrally positioned insert like the fletching shownin FIG. 118, defining essentially spiral fins or baffles or helicalprotrusions formed on the inside wall of the prechamber 12228 of thesingle burner head 12205 embodiment, just as shown in the multipleburner head embodiment of FIG. 119.

Following combustion in the combustion chamber 12250 the hot exhaustgases pass out of the combustion chamber 12250 and into a hot air, orgas, channel 12242 and into the heating chamber 12218 to heat the heatertubes 12209 of the heater heads 12205. The hot gas channel 12242 is alsoin communication with a cold air supply which communicates through asupplemental trim air passage 12240 extending through the top of theburner.

An important aspect of the single burner head embodiment is the abilityto control the combustion of all or part of the fuel in the prechamber12228 and the combustion chamber 12250 and provide an appropriatetemperature gas to the heater chambers 12218 based on this control ofthe combustion process. To this end there are initially two methods ofmanaging the combustion process in the single burner head design showndiagrammatically in FIGS. 123 and 124. The first method involvesproviding a higher relative percentage of fuel, i.e. a rich fuelmixture, which is ignited and burns in the combustion chamber butbecause there is not sufficient oxygen in such a rich mixture, theentire combustion process is not completed in the center combustionchamber and is to some extent cooler than if the combustion process wascompleted. The cooler partially un-combusted fuel is then circulated upinto the hot gas channel where the remaining fuel mixes with additionalair and can more fully complete the combustion process. This means thatthe gas in the center combustion chamber is not as hot where somepercentage of the fuel is not combusted, but that an additional desiredpercentage of fuel can be combusted in the secondary combustion processto control the temperature of the hot gases being provided into theheater chamber. In other words there may not be complete combustion inthe center combustion chamber so that the combusted gases are cooler,and the temperature can be raised if necessary by applying additionalair in the hot gas channel so that further combustion is accomplishedraising the gas temperature in the heater chambers. As shown in FIG.123, a rich fuel mixture is added at step 123101, ignition occurs atstep 123102, hot gases and uncombusted fuel are circulated to theheating chamber at step 103103, sensors measure temperatures incombustion and heater chambers at step 123104, and based on presetlevels air is added to increase combustion at step 123105.

The second method shown in FIG. 124 involves providing a lower relativepercentage of fuel, i.e. a normal fuel mixture which burns very hot inthe combustion chamber and as it exits, the air provided via the airtubes, also called trim air, is used to manage e.g. cool the temperatureof the hot gases entering into the heater chambers and provide some backpressure so that not as large a volume of hot combustion gases flowthrough a given hot gas channel into a specific heater chamber which isreceiving additional trim air. The benefit of a richer combustion isthat the cooler gases are less likely to damage the burner in the firstmethod whereas in the second complete combustion method, there may be adecrease in efficiency because of the cold air input into the heaterhead center. In the second method shown in FIG. 124, a fuel mixture isadded at step 124101, ignition occurs at step 124102, hot combustedgases are circulated to the heater chamber 124103, sensors measuretemperatures in combustion and heater chambers at step 124104, trim airis added to cool heater head and create back pressure in the heaterchamber at step 124105.

In either method, an important aspect of the invention is to maintain asequal a temperature across the four (4) heater heads 12203 with oneflame, and hence in the heater chamber 12218, as possible. Referringback to FIG. 122, the control of the gas temperature entering each ofthe four (4) separate heater chambers 12218 of the burner 12205 can beindividually controlled by directly controlling the temperature of thehot gas as in the second method, or controlling the combustion processas in the first method. The single burner is an important design becausethere is only a single flame, a single igniter, and few locations for apossible reliability issue.

Following heating of the heater heads 12203, a substantial amount of theheat not used to heat the working fluid remains in the exhaust gases andthus the efficiency of the entire engine can be increased by using theexhaust gas heat to preheat the incoming air. The hot combustion gasespass from the hot gas channel 12242 into the heater chambers and afterheating the heater heads and heater tubes therein, the hot combustiongases are forced out an exhaust inlet 12253 into the exhaust channel12244 which is defined by the exhaust manifold wall 12243. The exhaustpasses down along the exhaust channel 12244 exchanging a substantialamount of heat through the manifold wall and into the incoming cold airentering the preheater 12226 via the preheater channel 12241.

An insulation layer 12246 can be added around the entire burner betweenthe heater chambers 12242 and the preheater 12226 to keep heat fromtraveling out of the combustion chamber into the preheater 12226.

Restricting Flow

It is important to control the heat into each heater head to maximizethe efficiency and power of the engine by maintaining the heads at verysimilar temperatures. Moreover it is important to recognize that theengine is limited to a highest operating temperature by the heater headmaterial properties. In other words, the heater heads, or any particularheater head, cannot exceed the highest operating temperature. By way ofexample if one heater head is operating at the highest operatingtemperature, in the single burner embodiment described above the fuelflow cannot be increased to the engine to increase the temperature ofthe other heater heads. The hotter heater head must be cooled. Thecontrol of the flow of hot combustion gases past each heater head can becontrolled by the methods using the rich and lean combustion and trimair described above. Alternatively, another method of controlling thegas temperature applied to the heater heads shown in FIGS. 125A, 125Band 126, provides a flow of non-combustion gas (e.g. air) through an airintake 12536 into the bottom of each heater head adjacent the coolerplate 12525.

There is a non-combustion gas supply for each heater head as shown inFIG. 125A. An inlet 12536 supplies the non-combustion cooling gas to asupply line 12538 brazed to the outside of the burner base assembly12540 and includes an elbow 12542 to get around the corner whileoccupying minimal space. The supply line 12538 terminates at a flowdiverter ring 12544 which is located at the base of the heater head12503. The flow diverter ring 12544 has by way of example, twenty 0.100″diameter holes 12546 which create a restriction for any cooling gasessupplied to the heater head to eventually escape toward the exhaustoutlet (not shown) through the holes 12546. When the diverting air issupplied through the supply line 12538 to cool a heater head 12503, thehot combustion gases meet an increased resistance at that particularheater head and the remaining heads will experience a greater mass flowsince they provide less resistance to the incoming hot combustion gases.As a result, the amount of heat transfer—and therefore temperature—isdecreased for the particular heater head to which diverting air issupplied. The burner base assembly 12540 with cooling gas supplies lines12538 is shown in FIG. 125B.

Additional flow of air or other non-combustion gas, i.e. the restrictingflow, at the bottom of each heater head, although the flow can enter atother locations relative to the heater head as well, creates a backpressure in the heater chamber 12518 which restricts the amount of hotexhaust gases which can pass next to the heater head. The flow diverterring 12544 directs the cooling gas into the heater chamber andessentially merges the colder restricting gas or air with the hotexhaust gases attempting to exit the heater chamber 12518 into theexhaust. As discussed above, the cool non-combustion gas provides theadditional benefit of cooling the hot heater head in addition toproviding a back pressure which restricts the amount of hot combustiongases entering into the heater chamber 12518.

When restricting non-combustion gas, or air for example is applied inthis manner to the heater head 12503 which is too hot, the combustiongases meet an increased resistance at the particular head and theremaining heads will experience a greater flow of combustion gases sincethey provide less resistance. As a result, the one hot heater headdecreases in temperature and the other heads will increase intemperature, thereby decreasing the difference in temperature betweenthe heads and allowing an increase in fuel flow so a higher averagetemperature can be sustained. A suitable control system should allow avery close temperature tolerance between all the heads and providemaximum temperatures at each head to be essentially equal. As a result,efficiencies can be maximized.

A control scheme is shown in FIG. 126 for controlling the addition ofrestricting non-combustion gas to provide cooling and/or restrictiveflow to the heater chambers 12518. A flow controller FC is modulated tocontrol the hottest temperature heater head 1, 2, 3 and 4 to apredetermined target temperature. Valves 12669, in this caseservo-operated pinch valves, although other types of valves could beused as well, are opened between an air source, for example blower B oranother additional blower, and the heater chambers 12618 to allow theflow of cooling gas/air to the hottest of the heater heads 12603. As thecertain heater head 12603 and/or heater chamber 12618 temperaturesapproaches a desired average temperature, the flow of cool gas/air isreduced by closing the valves 12669.

The servo-operated pinch valves 12669 in this example, are controlled bythe flow controller FC for positioning of each servo between thefull-open and full-closed valve state. With this control scheme in anexemplary trial, the following important observations were made:

-   -   As diverting air is applied to a single head to reduce its        average temperature, the average temperature of each of the        remaining heads will increase in temperature by approximately ⅓        the reduction of the first head;    -   The full effect of the diverting air has a response of        approximately 2-3 minutes, however results are immediately        apparent;    -   Once temperatures reach the desired level, the amount of        diverting air applied typically should be reduced to maintain a        stable temperature reading.

FIG. 127 represents a snapshot of data collected in an initial trial ofthe control scheme and pinch valves 12669. The total average headtemperature increased by 43° C. from 893° C. to 936° C. whilemaintaining a maximum temperature of 980° C. Condition “B” representsthe data collected while diverting air was applied but the fuel flow wasunchanged. The end condition “C” was maintained with a total divertingmass flow of 1.45 g/s, or 15.4% of the total flow through the burner(7.9 g/s through the inlet of the burner for an exhaust oxygen level of˜7.3%). Fuel mass flow was increased from 0.303 g/s of Biodiesel to0.324 g/s.

Data was recorded for points of increased engine speed and crankcasepressure (2000 rpm/600 psi, 2500/600, 3000/600, 2000/750, 2500/750, and3000/750). For each set of conditions, the diverting air was changed tocontrol all max head temperatures as close to 980 C as possible, thusbringing the overall average head temperature to a maximum. Note that asengine speed increases, it appeared to become increasingly difficult tocool a single very hot head (i.e. the temperature spread between thehottest/coolest head becomes larger, even with max diverting airapplied).

FIG. 128 is a graphical representation of test results showing thebeneficial effects of the above described restricting air apparatus andcontrol scheme. The graph shows that for example when an appropriatediverting air flow was provided to at least heater head 2 (the hottesthead), the average heater head 2 temperature fell from 962 C to 949 C.And while heater head 4 remained substantially the same, heater heads 1and 3 raised their average temperatures from 829 C to 993 C and 860 C to926 C. The net effect brings the heater heads into much closertemperature tolerances, max and avg. as seen in the graph, and thusgreater engine efficiency and less likelihood of engine damaging fromover-heating certain heater heads.

The effect of the diverting air on providing even head temperatures andthe ability to possess individual head control in a single burner/fourheater head engine has demonstrated significant success especially whencompared to a four-burner/four heater head. With the above describedrestricting air apparatus and control scheme the additional hardwarerequired for diverting air is much cheaper (stainless steel and siliconetubing and four valves for supplying non-combustible gas/air); and therequired software controls are more or less the same. A single flamerequires only one igniter, fuel injector, and flame detector (and allowsfor direct control of the fuel/air ratio), making it easier and cheaperto assemble.

The current results provide approximately a 60° C. reduction intemperature on a single head without exceeding much more than 25% of thetotal mass flow through the burner. The result will also raise thetemperature of three cold heads by approximately 20° C., providing about80° C. in difference from the burner/engine steady state. To providethis diverting air, the blower will of course consume additional energyand reduce the power out of the engine. However, the desired net resultthat the 10-40 watts required blower power will be on the order of 20%of the additional power that is then able to be produced by the enginewith the greater efficiency.

While the principles of the invention have been described herein, it isto be understood by those skilled in the art that this description ismade only by way of example and not as a limitation as to the scope ofthe invention. Other embodiments are contemplated within the scope ofthe present invention in addition to the exemplary embodiments shown anddescribed herein. Modifications and substitutions by one of ordinaryskill in the art are considered to be within the scope of the presentinvention.

What is claimed is:
 1. An external combustion engine comprising: acrankcase, the crankcase comprising a crankcase gas and lubricating oil,the crankcase gas having a crankcase pressure; at least two heaterheads; at least two pistons, each of the at least two pistonreciprocating within one of the at least two heater heads; a workingvolume defined in part by the at least two heater heads and at least twopistons; at least one pair of piston rods comprising a first piston rodand a second piston rod, each piston rod coupled to one of the at leasttwo pistons and located partially in the working volume; a drivemechanism located in the crankcase, the drive mechanism comprising arotating crankshaft and a rocking beam, the crankshaft rotating therocking beam, the first piston rod attached to a proximal end of therocking beam and the second piston rod attached to a distal end of therocking beam, the rocking beam constrains the first piston rod to move180 degrees out of phase with the second piston rod; an airlockseparating the crankcase and the working volume, the airlock comprisingan airlock gas, the airlock gas having an airlock pressure, wherein theat least one pair of piston rods pass through the airlock and eachpiston rod is sealed to the airlock by a rolling bellows seal; and aairlock pressure regulator comprising at least one valve and acompressor, the airlock pressure regulator maintaining a pressuredifferential between the crankcase pressure and the airlock pressure. 2.The external combustion engine of claim 1 wherein the airlock pressureregulator further comprises an oil filter fluidly connected to thecrankcase, the compressor being fluidly connected to the filter and thevalve being a pressure regulating spool valve fluidly connected to atleast one of the filter, the compressor and the airlock.
 3. The externalcombustion engine of claim 2 wherein a position of the pressureregulating spool valve is responsive to differences between the airlockpressure and the crankcase pressure.
 4. The external combustion engineof claim 3, further comprising a controller wherein the controlleroperates the compressor based on the position of the spool valve.
 5. Theexternal combustion engine of claim 3 wherein the position of thepressure regulating spool valve is determined with a proximity sensor.6. The external combustion engine of claim 3 wherein the pressureregulating spool valve further comprises an under-pressure relief valvethat limits the airlock pressure to less than a predetermined valueabove the crankcase pressure.
 7. The external combustion engine of claim1, wherein the airlock pressure regulator controls the airlock pressureto be not less than a first predetermined value and not more than asecond predetermined value above the crankcase pressure.
 8. The externalcombustion engine of claim 1, wherein the airlock pressure regulatorcontrols the airlock pressure to be not more than a first predeterminedvalue below the crankcase pressure and not more than secondpredetermined value above the crankcase pressure, while the drivemechanism is not in motion.
 9. The external combustion engine of claim1, further comprising a controller, wherein the controller prevents thedrive mechanism from being in motion while the airlock pressure is lessthan a predetermined value above the crankcase pressure.
 10. Theexternal combustion engine of claim 1 further comprising a pressureequalization orifice between the airlock and the working volume, wherebythe air lock pressure is approximately equal to an average of a workingvolume pressure.
 11. A reciprocating apparatus with a pressure regulatorfor maintaining a pressure differential between a first closed volumeand a second closed volume, the first closed volume filled with a gasand liquid, the gas in the first closed volume having a first pressure,the second closed volume filled with the gas, the gas in the secondclosed volume having a second pressure, the reciprocating apparatus witha pressure regulator comprising: a boundary between the first closedvolume and the second closed volume, the boundary comprises at least oneflexible section, the flexible section attached to a reciprocating rod,the reciprocating rod driven by a rotating shaft via a rocking beam inthe first closed volume; and the pressure regulator comprising: a filterfluidly connected to the first closed volume; a compressor having aninput and an output, the input fluidly connected to the filter; and abidirectional pressure regulating spool valve having fluidic connectionsto at least one of the filter, the compressor and the second closedvolume, wherein the fluidic connections are based on a position of thepressure regulating spool valve, wherein the bidirectional pressureregulator spool valve controls the second pressure to be not less than afirst predetermined value above the first pressure and not more than asecond predetermined value above the first pressure, while the rotatingshaft is turning.
 12. A reciprocating apparatus with a pressureregulator for maintaining a pressure differential between a first closedvolume and a second closed volume, the first closed volume filled with agas and liquid, the gas in the first closed volume having a firstpressure, the second closed volume filled with the gas, the gas in thesecond closed volume having a second pressure, the reciprocatingapparatus with a pressure regulator comprising: a boundary between thefirst closed volume and the second closed volume, the boundary includesat least one flexible section, the flexible section attached to areciprocating rod, the reciprocating rod driven by a rotating shaft viarocking beam in the first closed volume; the pressure regulatorcomprising: a filter fluidly connected to the first closed volume; acompressor having an input and an output, the input fluidly connected tothe filter; and a bidirectional pressure regulating spool valve havingfluidic connections to at least one of the filter, the compressor andthe second closed volume, wherein the fluidic connections are based on aposition of the bidirectional pressure regulating spool valve; and acontroller, wherein the controller prevents the rotating shaft fromturning while the second pressure is less than a predetermined valueabove the first pressure.